Control apparatus for drive system

ABSTRACT

A drive system composed of an engine and a transmission is controlled in accordance with a desired wheel toque corresponding to a position of an accelerator, and a present vehicle speed in such a way that a speed ratio of the transmission is determined in consideration with torque factors such as an air-fuel ratio on the engine side, thereby it possible to optimize the control in order to reduce the emission of exhaust substance such as NOx and to enhance the acceleration performance and the fuel economy.

BACKGROUND OF THE INVENTION

[0001] The present invention relates to a control apparatus for a drivesystem composed of an engine and a transmission, and in particular to acontrol apparatus for a drive system composed of an engine such as agasoline engine or a Diesel-engine and a transmission such as anautomatic transmission in combination of a torque converter and a geartrain, or a belt-and-pulley type continuous variable transmission.

RELATED ART

[0002] For example, Japanese Patent Publication No. 63-45977 andJapanese Patent Publication No. 63-45976 disclose therein a controlapparatus which sets a desired wheel torque (which corresponds to atorque that can obtained at the final speed change gear such asdifferential gear or a speed ratio between the engine speed and thewheel speed since the engine torque is substantially constant,irrespective of an engine speed, over an operating speed range) inaccordance with an opening degree of the throttle valve of an enginewhich is manipulated by a driver. Further, it has been well-known that acontrol apparatus which sets the speed ratio in accordance with a torqueof an engine.

[0003] In a conventional control apparatus for a drive system composedof an engine and a transmission, a desired wheel torque has been set inaccordance with only an engine torque, irrespective of an air-fuel ratio(a fuel quantity in an engine cylinder), an intake valve closing angle,a supercharging pressure and a ratio between working and compressionstrokes, which are parameters for the engine torque. That is, since thedesired torque has been directly set in relation to an engine torque anda vehicle speed, the consistence between the fuel economy and theacceleration performance has been difficult. Accordingly, in the case ofthe speed-up of a vehicle by changing the torque of an engine or that ofa transmission, depending upon a taste of a driver (that is, whether heis fond of high acceleration or not) or a recognition of environmentaround the vehicle, the following problems have been raised, that is,should the acceleration performance be heighten while the torque of thetransmission is maintained to be low with respect to an engine torque,the fuel economy would deteriorate. On the contrary, should the fueleconomy be enhanced by increasing the torque of the transmission withrespect to an engine torque, the acceleration performance woulddeteriorate.

[0004] Further, an in-cylinder fuel injection engine has been preferablyused as an engine constituting the drive system in order to preformprecise and complicated control. However, a conventional controlapparatus has been adapted to control the timing of fuel injection andthe timing of ignition under such a condition that the air volume is setto be constant. Accordingly, should an in-cylinder fuel injection enginehaving a ratio between working and compression strokes of less than 1 becontrolled, the mixture would be locally overrich, causing generation ofsoot in the case of a large fuel injection volume, or the mixture wouldbe excessively lean around a spark plug so as to cause the combustionunstable in the case of a small fuel volume. Further, if the fuel volumeincreases under such a condition that the air volume is constant, theair-fuel ratio decreases, causing increasing of nitrogen oxide (NOx)emission.

[0005] For example, an in-cylinder fuel injection engine disclosed inJapanese Laid-Open Patent No. 60-30420, incorporates a fuel injectionvalve directed to a spark plug, and an air injection value adapted toinject air which interferes with fuel injected from the injection valve,and accordingly, in the case of a less fuel induction volume during lowload operation, air is injected into fuel jetted from the fuel injectionvalve to the spark plug so as to concentrate the fuel around the sparkplug. In this arrangement, lean-burn operation and reduction of pumpingloss can be carried out. However, since such an in-cylinder injectionengine additionally requires the above-mentioned air-injection valve,not only the manufacturing cost of the engine is increased, but alsounburnt hydrocarbon emission cannot be sufficiently reduced even withthe provision of the above-mentioned arrangement.

SUMMARY OF THE INVENTION

[0006] The present invention is devised in view of the above-mentionedproblems inherent to a conventional control apparatus for a drive systemcomposed of an engine and a transmission, and accordingly, a firstobject of the present invention is to provide a control apparatus for adrive system composed of an engine and a transmission, which can performflexible control so as to enhance both fuel economy and accelerationperformance.

[0007] To the end, according to a first aspect of the present invention,a control apparatus for a drive system composed of an engine and atransmission is provided with a computing means for controlling thetorque of the transmission and the air-fuel ratio of the engine inaccordance with a desired wheel torque and a vehicle speed.

[0008] According to a specific form of the present invention, a controlapparatus for a drive system is provided with a computing means forcontrolling the torque of a transmission and the closing angle of anintake valve in relation to each other in accordance with a desiredtorque and a vehicle speed. According to another specific form of thepresent invention, a control apparatus for a drive system is providedwith a computing means for controlling the torque of a transmission anda supercharging pressure in relation to each other in accordance with adesired wheel torque and a vehicle speed, or a computing means forcontrolling the torque of a transmission and a ratio between working andcompression strokes in relation to each other in accordance with adesired wheel torque and a vehicle speed.

[0009] The above-mentioned computing means carries out control operationin such a way that the fuel consumption and the acceleration performanceare optimized with the use of a performance chart in accordance with adriver's taste or an operating environment of a vehicle.

[0010] Further, a second object of the present invention is to provide acontrol apparatus for controlling an engine used in a drive system of avehicle, preferably for controlling an in-cylinder fuel injection enginewhose ratio between working and compression strokes can be set to beless than 1, which can prevent generation of soot, unstable combustion,and increasing of NOx emission.

[0011] To the end, according to the present invention, the controlapparatus for an internal combustion engine comprises a fuel injectiontiming control means for controlling the volume of fuel injected from ajet port of a fuel injection device which is located in a combustionchamber of the engine, and the fuel injection timing, and an intakevalve opening and closing control means for controlling the opening andclosing timing of an intake valve of the engine, the fuel injectiontiming control means controlling the volume of the fuel and the fuelinjection timing in accordance with a variation in air volume to beburnt in the combustion chamber.

[0012] Further, a third object of the present invention is to provide apreferable fuel injection valve for the above-mentioned engine, whichcan reduce the manufacturing cost of the engine, which can reduce theemission of unburnt hydrocarbon as far as possible, and which canexhibit stable combustion over a wide operating range of the engine.

[0013] To the end, according to the present invention, the fuelinjection valve comprises a means for injecting fuel in a decreasedinjection splay angle, and a means for injecting fuel in an increasedinjection splay angle.

[0014] Specific explanation will be hereinbelow made of specificembodiment forms of the present invention with reference to theaccompanying drawings in which:

BRIEF DESCRIPTION OF THE INVENTION

[0015]FIG. 1 is a schematic view illustrating a drive system appliedtherein with a control apparatus according to the present invention;

[0016]FIG. 2 is a control block diagram illustrating a computingapparatus for controlling a speed ratio and an engine torque,incorporated in a drive system shown in FIG. 1;

[0017]FIG. 3 is a control block diagram illustrating an air-fuel ratiocomputing means incorporated in the control apparatus shown in FIG. 1,

[0018]FIG. 4 is a control block chart illustrating a first concept ofthe control apparatus shown in FIG. 1;

[0019]FIG. 5 is a control block diagram illustrating a drive torquecomputing means incorporated in the control apparatus shown in FIG. 1;

[0020]FIG. 6 is a control block diagram illustrating a first variantform of the computing apparatus;

[0021]FIG. 7 is a control block diagram illustrating a second variantform of the computing means;

[0022]FIG. 8 is a performance chart;

[0023]FIG. 9 is a sectioned performance chart;

[0024]FIG. 10 is a detailed view of the performance chart shown in FIG.9;

[0025]FIG. 11 is a detailed view of the performance chart shown in FIG.10;

[0026]FIG. 12 is a performance chart relating to vehicle speed v.s.engine speed (speed ratio);

[0027]FIG. 13 is a performance chart for explaining variation in theair-ratio with respect to vehicle speed:

[0028]FIG. 14 is a performance chart;

[0029]FIG. 15 is a wheel torque v.s. engine torque performance chart;

[0030]FIG. 16 is an F-V performance chart;

[0031]FIG. 17 is a view illustrating a first embodiment of an engineused in the drive system shown FIG. 1;

[0032]FIG. 18 is a view for explaining relationship between position ofan acceleration pedal and fuel volume;

[0033]FIG. 19 is a view for explaining relationship among the fuelvolume, injection timing and ignition timing;

[0034]FIG. 20 is a view for explaining relationship between operation ofan opening and closing cam for an intake valve and crank angle;

[0035]FIG. 21 is a view for explaining relationship among fuel volume,air volume and intake valve closing timing;

[0036]FIG. 22 is a view for explaining relationship between fuel volumeand air-fuel ratio;

[0037]FIG. 23 is a view for explaining relationship between fuel volumeand emission of nitrogen oxide and hydrocarbon;

[0038]FIG. 24 is a view for explaining relationship among fuel volume,fuel consumption rate and ratio between working and compression strokes;

[0039]FIG. 25 is a view for explaining relationship between fuel volumeand fuel consumption rate;

[0040]FIG. 26 is a flow chart for explaining control of the firstembodiment of the present invention;

[0041]FIG. 27 is a schematic view illustrating a control apparatus forair volume;

[0042]FIG. 28 is a view for explaining relationship among air-fuelratio, torque and emission of nitrogen oxide;

[0043]FIG. 29 is a view for explaining relationship among fuel volume,air-fuel ratio and air volume;

[0044]FIG. 30 is a view for explaining relationship among air-fuelratio, quantity of nitrogen oxide and air-volume;

[0045]FIG. 31 is a schematic view illustrating a second embodiment of anengine used in the drive system shown in FIG. 1;

[0046]FIG. 32 is a schematic view showing a part around the cylinderhead of the engine shown in FIG. 31;

[0047]FIG. 33 is a graph for explaining relationship between radius ofcurvature of an intake pipe and pressure loss;

[0048]FIG. 34 is a schematic view for illustrating an intake and exhaustvalve drive system show in FIG. 1;

[0049]FIG. 35 is a graph for explaining stability of engine speed causedby intake air port injection and in-cylinder fuel injection duringidling operation:

[0050]FIG. 36 is a sectional view illustrating a first embodiment of afuel injection valve (in a condition of wide angle atomization);

[0051]FIG. 37 is a sectional view illustrating the fuel injection valveshown in FIG. 36 (in a condition of narrow angle atomization);

[0052]FIG. 38 is a view for explaining positional relationship betweenthe fuel injection valve shown in FIG. 36 and a spark plug;

[0053]FIG. 39 is a graph for explaining relationship between injectiontiming and density of hydrocarbon for different divergent atomizationangles in the case of using the fuel injection valve shown in FIG. 36;

[0054]FIG. 40 is a graph for explaining relationship between density ofhydrocarbon and rate of catalyst conversion from nitrogen oxide intonitrogen by catalyst:

[0055]FIG. 41 is a sectional view illustrating a first embodiment of afuel injection distributor (which feeds fuel in a #1 engine cylinder)shown in FIG. 31;

[0056]FIG. 42 is a sectional view illustrating the fuel injectiondistributor (which feeds fuel into a #4 engine cylinder) shown in FIG.41;

[0057]FIG. 43 is a sectional view along line XIII-XIII shown in FIG. 41;

[0058]FIG. 44 is a timing chart for explaining the timing of fuel supplyinto an engine cylinder by the fuel distributor shown in FIG. 41;

[0059]FIG. 45 is a view for explaining the structure of a catalystconverter shown in FIG. 31;

[0060]FIG. 46 is a graph showing variation in the function, caused byvariation in temperature, of the catalyst converter shown in FIG. 45;

[0061]FIGS. 47A to 47G are views for explaining operation of the Millercycle engine shown in FIG. 31;

[0062]FIG. 48 is a sectional view illustrating a first variant form ofthe fuel injection valve (in a closed condition) shown in FIG. 31,

[0063]FIG. 49 is a sectional view illustrating the fuel injection valve(in a condition of wide angle atomization) shown in FIG. 31;

[0064]FIG. 50 is a sectional view illustrating the fuel injection valve(in a condition of narrow angle atomization) shown in FIG. 49;

[0065]FIG. 51 is a graph for explaining relationship between excess airfactor and density of hydrocarbon for different atomization conditions;

[0066]FIG. 52 is a graph for explaining relationship between fuelinjection timing and density of hydrocarbon in a rod-like atomizationcondition;

[0067]FIG. 53 is a sectional view illustrating a second variant form thefuel injection valve shown in FIG. 31;

[0068]FIG. 54 is a sectional view illustrating essential part of thesecond variant form of the fuel injection valve shown in FIG. 53;

[0069]FIG. 55 is a sectional view illustrating a third variant form ofthe fuel injection valve show in FIG. 51;

[0070]FIG. 56 is a circuit diagram illustrating a solenoid drive circuitin the fuel injection valve FIG. 55;

[0071]FIG. 57 is a view for explaining positional relationship between afuel injection valve and a spark plug:

[0072]FIG. 58 is a timing chart for explaining fuel injection timing;

[0073]FIG. 59 is a sectional view for illustrating a second modifiedform of the distributor shown in FIG. 31;

[0074]FIG. 60 is a sectional view for illustrating a third modified formof the distributor shown in FIG. 31;

[0075]FIG. 61 is a schematic view for illustrating an embodiment of afuel pump shown in FIG. 31;

[0076]FIG. 62 is a detailed view showing a part of the fuel pump asviewed in the direction of the arrow XXXII in FIG. 61;

[0077]FIG. 63 is a timing chart for illustrating fuel injection timingin the case of using the fuel pump as shown in FIG. 62.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0078] A continuous variable transmission 3 is used as an example of atransmission of a drive system in which a control apparatus according tothe present invention is applied, in the explanation which will behereinbelow made. However, it should be noted that the continuousvariable transmission 3 is not an indispensable matter of the presentinvention, but any other transmission such as a torque converter or aspeed change gear can be similarly used, instead of the continuousvariable transmission, in the present invention. Further, a gasolinetype in-cylinder fuel injection engine is used in an example of anengine in the drive system according to the present invention, but itshould be noted that a gasoline type intake port fuel injection engineor a Diesel-engine can be also used, instead of the gasoline typein-cylinder fuel injection engine, in the present invention.

[0079] A drive system shown in FIG. 1 is composed of an engine 1 and atransmission 3 which is coupled to the engine 1 through the intermediaryof a clutch 2, and accordingly, an output is transmitted to a wheel 4from the engine 1 through the clutch 2 and the transmission 3. Further,the speed ratio of the transmission 3 is controlled by a transmissioncontrol means 22 while the clutch 2 is controlled by a clutch controlmeans 2, and the speed ratio of a transmission 15 for a compressor 7 iscontrolled by a speed ratio control device 16.

[0080] An air filter 5, an air-flow sensor 6, the compressor 7, anintake pipe 8 and an intake valve 9 are provided on the intake side ofthe engine 1, and accordingly, intake-air is fed into an engine cylinder10 therethrough, successively. The compressor 7 is driven by the engine1 through the intermediary of the transmission 5. The opening andclosing timing of the intake valve 9 is controlled by a valve openingand closing timing control device 17 which is driven by a cam shaft 18.

[0081] Further, fuel is injected into the engine cylinder 10 by a fuelinjection valve 11, and a mixture of an air and fuel in the enginecylinder 10 is ignited and burnt by a spark plug 12. Exhaust gas isemitted into the atmosphere from the cylinder 10 through an exhaust pipe13 and a catalytic converter 14. An air-fuel ratio sensor 19 is attachedto the exhaust pipe 13, and delivers its output to a control unit 20.

[0082] The speed ratio control valve 16, the clutch control means andthe transmission control means 22 are controlled by the control unit 23which is connected to a control unit 20 through the intermediary of alocal area network (LAN) 24. An engine speed is detected by an enginespeed sensor 25 which delivers an output signal 23. A position of anaccelerator pedal 26 is detected by a potentiometer 27 and is deliveredto the control unit 23 which is connected thereto with a vehicleenvironment recognition means 28 and a driver recognition means 29.

[0083] A vehicle front monitor radar 30 delvers signal to theenvironment recognition means 28 which therefore recognizes a runningenvironment, that is, whether an obstacle to running, such as a vehicle,is present in front of the instant vehicle or not is determined. Thepotentiometer 27 delivers its output to the recognition means 29 whichtherefore recognizes whether the driver is in favor of a sporty drivepattern or not in view of a time-variation in the output of thepotentiometer 27. If the time-variation is large, it does means that theposition of the accelerator pedal 26 is abruptly changed, andaccordingly it is recognized that the driver is in favor of a sportydrive pattern. Further, the control unit 23 obtains a desired wheeltorque from a position of the accelerator pedal 26, delivered from thepotentiometer 26. A vehicle speed sensor 31 delivers its output signalto the control unit 23.

[0084] The control unit 23 incorporates a speed ratio and engine torquecomputing means 100 for computing a speed ratio of the transmission 3,an engine torque and a slip of the clutch 2 from the desired wheeltorque and the vehicle speed. Meanwhile the control unit 20 incorporatesan air-fuel ratio computing means 101 as shown in FIG. 3 for computing amixture air-fuel ratio from the engine torque. If the desired wheeltorque is large, the air-fuel ratio is decreased so as to enrich themixture in order to increase the output torque of the engine. That is,the speed ratio and engine torque computing means 100 and the air-fuelratio computing means 101 are associated with each other. Accordingly,as the speed ratio between wheel speed and engine speed is large, thedesired torque is set to become less, and accordingly, the air-fuelratio decreases. Meanwhile, if the speed ratio is set to be small, thedesired wheel torque decreases, but the air-fuel ratio increases. Inview of this point, it is noted that the relationship between the speedratio and the torque has been manually changed over by means of apower/economy selection lever in prior art, and has not be associatedwith the air-fuel ratio computing means 101, and accordingly, the priorart device is inferior in flexibility.

[0085] Referring to FIG. 4 which is a block diagram for facilitating theunderstanding of the concept of the present invention, a speed ratio, aslip of the clutch and an air-fuel ratio are computed from a desiredtorque and a vehicle speed. Since the in-cylinder fuel injection engine1 is used, the charged air volume is constant, and accordingly, theair-fuel ratio is controlled by adjusting the volume of fuel injectedfrom the fuel injection valve 11. Further, the clutch slip control means21 controls the slipping of the clutch, and further, the transmissioncontrol means 22 controls the speed ratio. As mentioned above, thecomputing means 102 is in the combination of the speed ratio and enginetorque computing means 100 and the air-fuel ratio computing means 101.

[0086] The control unit 20 incorporates a power torque computing means102 which computes a torque generated at the engine from a desired wheeltorque and a rate of variation in engine speed, and computes the volumeof fuel injected from the fuel injection valve 11 so as to obtain thegenerated torque as shown in FIG. 5. Accordingly, a decrease in torqueused for increasing the engine speed can be compensated. Conventionally,the relationship between the air-fuel ratio and the fuel volume has beenfixed so that the above-mentioned compensation has been impossible.Further, it is noted that the position (degree of depression) of theaccelerator pedal can be alternatively used, instead of the rate ofvariation in engine speed.

[0087] By the way, as already well-known, if the engine speed isincreased before acceleration, the acceleration performance can beenhanced. However, this causes an increase in fuel consumption. As shownin FIG. 6, an output from the radar is delivered to the environmentrecognition means 28 which therefore computes a maximum desired wheeltorque and delivers the same to the computing means 102. If no obstacleis present in front of the vehicle so that the maximum desired torque islarge, the speed ratio is set to a small value. Meanwhile, if anobstacle is present in front of the vehicle, or a traffic jam occurs,the speed ratio is increased so as to save fuel consumption. Further, ifthe driver recognition means 29 determines that the driver is in favorof a sporty operating pattern, the maximum desired wheel toque becomeslarge, and accordingly, the speed ratio is set to be small.

[0088] As shown in FIG. 7, if the desired torque of the engine is small,the valve opening and closing timing control device 17 retards theclosing timing of the intake valve so as to reduce the compression work,and accordingly, the fuel economy can be enhanced. At this time, if theair-fuel ratio is controlled to a stoichiometric ratio, three-waycatalyst can be used for the catalytic converter 14. Further, if thedesired torque of the engine is small, the speed ratio control device 16lowers the revolution speed of the compressor 7 so as to reduce thesupercharging pressure down to a value nearly equal to the atmosphericpressure. Accordingly, the compression work of the compressor 7 isdecreased so as to enhance the fuel economy. If the desired torque islarge, the revolution speed of the compressor 7 is increased so as toincrease the supercharging pressure while the air-fuel ratio ismaintained to be stoichiometric. Alternatively, the closing timing ofthe intake valve is advanced to increase the charged air volume of thecylinder 1.

[0089] Referring to FIG. 7, the speed ratio and engine torque computingmeans 100 obtains an engine desired torque and a speed ratio, anddelivers the engine desired torque to the speed ratio control device 16and the valve opening and closing timing control device 17 so as tooptimumly control the revolution speed of the compressor 7 and theclosing angle of the intake valve 9.

[0090] Next, explanation will be made hereinbelow made of the operationof this embodiment of the invention.

[0091] Referring to FIG. 8, denoting that the ratio between wheel andengine speeds as a speed ratio x, if x is large, the wheel torque fallsinside of a curve A while if x is small, the wheel torque falls insideof a curve B. If a desired value Fo is given to the wheel torque F withrespect to the vehicle speed, x in the curve B is selected if F₁>F₀ andV₁<V₀. If the continuously variable transmission is used, the point B₂continuously varies to a point A₁. At this time, the speed of the engineis highest. Further, the point B₁ continuously varies to the point A₁.At this time the engine speed is lowest.

[0092] In view of the fuel economy, the engine is preferably driven at aspeed and a load which are as low and high as possible, respectively.Accordingly, it is driven along a curve A₁-B₁. If F=F₂, it is drivenalong a curve A₁-A₂. If F=F₁, it is driven along a curve B₁-B₂. If V>V₁and F<F₁, the output power of the engine is controlled at a speed ratioB. If F<F₁ and V=V₁, it is driven along a curve V₁-B₁ while if V=V₂, itis driven along a curve V₂-A₁. If V<V₂, the speed of the engine islowered if no slip occurs, and accordingly, the clutch is slipped inorder to maintain the speed of the engine. A curve RL shown in thisfigure indicates a desired torque F during a horizontal road surface,with which the engine cannot be drive above a point P₂ with a leanmixture.

[0093] In order to enlarge the vehicle speed range and the wheel torquerange, the mixture is enriched. Accordingly, the torque F₂ is increasedup to a value F₂′ with which the vehicle can ascend a slope. Further,the torque F₁ is increased up to a value F₁ so that the driveable pointis increased to a value P₂′. Referring to FIG. 9, in a range a, it isdriven at a lowest engine speed with a partial load and a lean mixture.In a range β where the speed ration is lowest, it is driven at a highestengine speed with a rich mixture. In a range γ, it is driven at ahighest speed ratio x with a rich mixture. In a range ε where the speedratio x is highest, it is driven with a rich mixture. In a range ε, itis driven with a lean mixture while the throttle valve is fully opened.The speed ratio x varies, depending upon the speed V.

[0094] The range ε can be subdivided as shown in FIG. 10. In the rangeε₁, it can also be driven at a highest ratio x with a rich mixture.However, at this time, since the fuel consumption is high in thiscondition in comparison with the condition in which the engine is drivenwith a lean mixture while the throttle valve is fully opened, it isdriven in the latter condition.

[0095] In the range ε₂, it can be driven at a low engine speed with arich mixture. Accordingly, on the left side of a curve E₁-E₂, it isdriven with a rich mixture while the throttle valve is fully opened inorder to restrain the engine speed from increasing. In the range ε₃, itis driven with a rich mixture and a fully opened throttle valve. On theleft side of the curve E₁-E₂, it is driven with a rich mixture and afully opened throttle valve. That is, though it can be driven with alean mixture and a fully opened throttle valve in the range ε₂, themixture is set to be rich in a range where the engine speed exceeds, forexample, 3,000 rpm so as to lower the engine speed since the fuelconsumption is increased due to mechanical friction as the engine speedis increased.

[0096] As shown in FIG. 11, in the ranges α, β, the air-fuel ratio iscontrolled in the range from 25 to 80, that is, the higher the air-fuelratio, the larger the fuel volume. In a range δ, the air-fuel ratio iscontrolled in a range from 25 to 12. In the range ε, on the left side ofthe curve E₁-E₂, the engine is driven at an air-fuel ratio of 25 whilethe wheel torque F is adjusted, depending upon a speed ratio x. In therange γ, the air-fuel ratio is set to 12 along a curve a₁-a₂ while it isset to 25 along a curve b₁-b₂. In this condition, the engine speed ishighest, that is, 6,000 rpm.

[0097] Along a curve C shown in FIG. 11, the air-fuel ratio is set to 12so as to restrain the engine speed below 3,000 rpm. Accordingly, theair-fuel ratio is set to be low in order to lower the engine speed,rather than the air-fuel ratio is set to 25 so as to increase the enginespeed. In a range between the curves E, C, the air-fuel ratio is changedfrom 25 to 12 so as to maintain the engine speed at 3,000 rpm.

[0098] Referring to FIG. 12, under the load RL during running on ahorizontal road, at first, the engine speed is low, that is, 1,000 rpm,and the vehicle speed is increased by increasing the speed ratio. Whenthe speed ratio x reaches its maximum value X_(max), the vehicle speedis increased by increasing the engine speed. When the vehicle speedexceeds a value V₃ at which the engine speed reaches 3,000 rpm, thespeed ratio is decreased while the torque F is increased, and further,the engine speed is increased so as to increase the vehicle speed up toa value V₄ as shown by the broken line in FIG. 12. At V₄, the enginespeed reaches 6,000 rpm. At this time, the vehicle speed reaches at avalue V₅ by increasing the speed ratio. If the air-fuel ratio isdecreased at the point V₃, the vehicle speed is increased as shown bythe solid line in FIG. 12 without decreasing the speed ratio x.

[0099] Referring to FIG. 13 which shows variation in the air-fuel ratiowith respect to the vehicle speed, the nitrogen oxide emission increasesaround an air-fuel ratio of 16. This point can be shifted to a highspeed side as shown by the broken line.

[0100] If the vehicle speed is increased from the point P on the curveRL, when the degree of depression of the accelerator pedal is large,that is, the position thereof is deep, the air-fuel ratio is set from 25to 12 in the range 8 or in the range γ. If the degree of depression ofthe accelerator pedal is small, the air-fuel ratio is set from 25 to 80.Upon shifting into the range δ, the speed ration x is decreased, andaccordingly, the engine speed should be increased so that a part of thetorque generated by the engine is consumed to increase the engine speed,and accordingly, the acceleration performance correspondingly lowers.For the compensation therefor, an extra torque is generated. That is,the wheel torque F is exhibited by the following expression:

F=k(n _(e) /V)T=k ₁(T/x)  EX1

[0101] where k, k₁ are constants, T is an input torque to thetransmission, n_(e) is an engine speed and V is a vehicle speed. If theengine generated toque is denoted as T_(e) which is given as follows:

T _(e) =T+I*dn _(e) /dt  EX2

[0102] where I is an inertia term of a movable part.

[0103] By substituting EX1 into EX2 while setting as F=F₀, the followingexpression can be obtained;

T _(e) =F ₀ *x/k ₁ +I*dn _(e) /dt  EX3

[0104] Accordingly, if the engine torque is increased from thesteady-state engine torque T_(e) by a value given by the second term onthe right side of the expression.

[0105] In FIG. 14, in a range of A₁-A₁′-B₁′-B₁, if the engine is drivenby shifting the air-fuel ratio from 25 to 12, the vehicle can beaccelerated without changing the speed ratio x. Further, in the range α,if the minimum value of the engine speed has been increased beforehand,the accelerating performance can be enhanced. However, an increase infuel consumption is inevitable. At a vehicle speed V₁, the speed ratio xcan be set to be either maximum or minimum. If the speed ratio x is setto be minimum, the wheel torque can be increased along the curve A₁′-A₂′without increasing the engine speed. On the contrary, if the speed ratiox is set to be maximum, the wheel torque along the curve B₁-B₂ can onlybe obtained if the engine speed is maintained to be constant. In orderto obtain a higher torque than that, it is required that the speed ratiox is decreased and the engine speed is increased. At this time, a partof the engine generated torque is consumed for accelerating the engineitself.

[0106] In the above-mentioned arrangement, estimating that the desiredwheel torque for acceleration is known before acceleration, the speedratio is set to be small beforehand if the desired wheel torque islarge, and accordingly, it is possible to prevent the accelerationperformance from lowering. However, the fuel consumption is slightlyincreased during steady-state operation. The desired wheel torquebecomes large when no vehicle as an obstacle is present in front, orwhen acceleration for passing, ascent of a steep slope or the like iscarried out. Such a condition can be known by detecting an environmentsurrounding the vehicle. It has been well-known that the presence of anobstacle in front can be detected by using a radar, a laser or an imageprocess. Further it has been well-known to detect a slope by using aninclination sensor or in accordance with a variation in runningresistance. Further, one driver is in favor of abrupt acceleration butanother driver in favor of fuel economy. It has been also well-knownthat this can be known by processing an accelerator pedal depressingpattern. If the driver in favor of a sport drive pattern, the speedratio x is set to be small during steady-state operation.

[0107] If the vehicle speed exceed the value V₄, the speed ratio xcannot be set to a minimum value, and accordingly, the wheel torquecannot exceeds a curve A₂′-B₂′. Also at this time, if the engine speedduring steady-state operation is set to be low, the torque duringacceleration is once lowered. The torque along A₁-A₂-B₂ can be attainedeven with a lean mixture. When the curve is exceeded, the mixture isenriched. At a vehicle speed V₅, the speed ration x is maximum, andaccordingly, the torque level is limited below the curve B₁-B₂. In orderto obtain the torque larger than that, it is required that the speedratio x is decreased while the engine speed is increased, or the speedratio x is set to be maximum while the mixture is enriched. In view ofthe acceleration performance, the latter is advantageous, but in view ofthe fuel economy, the former is advantageous. Accordingly, whether themixture is enriched so as to increase the torque or the speed ratio x isdecreased so as to increase the torque, is depend upon a driver's tasteor an environment around the vehicle.

[0108] In a Miller cycle engine in which the ratio between working andcompression strokes is equal to or less than 1, that is,working/compression stroke ratio≧1, the closing angle of an intake valveis adjusted so as to control the engine generated torque. Alternatively,the pressure of a supercharger is controlled so as to control thegenerated torque. If the closing angle is retarded, the compressionstroke is decreased so as to increase the above-mentioned stroke ratio,and accordingly, the expansion energy can be effectively used so as toenhance the fuel economy. In order to increase the engine generatedtorque, it is required to increase the compression stroke. However, thefuel economy is accordingly lowered. Whether the speed ratio x isdecreased so as to increase the wheel torque or the compression strokeis increased so as to increase the torque, depends upon the driver'staste or an environment around the vehicle at that time. Referring toFIG. 15, when the vehicle speed is constant, if the speed ratio x islarge, the engine torque is large, but if the speed ratio x is small,the engine torque is small. Up to a value T_(e1), the engine can bedriven, maintaining the stroke ratio constant, but it is required thatthe compression stroke is decreased upon shifting from T_(e1) to T_(e2).If the speed ration x is decreased, the engine can be driven,maintaining the stroke ratio constant in a range from F₀₁ to F₀₂ sincethe engine torque is less than the value T_(e1). However, if the speedratio x is excessively decreased, the engine speed is increased so asincrease the fuel consumption. If the engine speed exceeds 3,000 rpm,the speed ratio is set to be large. If acceleration is desired, thespeed ratio x is decreased, and if the fuel economy is essential, thevalue T_(e1) is set be as small as possible at an engine speed of 3,000rpm.

[0109] In the Miller cycle engine, in order to increase the enginetorque while maintaining the working and compression stroke ratio large,it is required to increase the supercharge pressure. Even though thecompression stroke is small, a large volume is charged into the enginecylinder so as to increase the engine torque. However, since thecompression work is increased. the speed ratio x is increased so as todecrease the desired engine torque until the engine speed becomes 3,000rpm.

[0110] Setting T=T_(e) in EX1, the following expression can be obtained:

F=k(n _(e) /V)*T _(e)  EX4

[0111] In a range A in FIG. 16, if it is operated with T_(e) =T _(e1),it is required to set the engine speed at a value higher than 3,000 rpm.In this phase, the engine torque T_(e) is increased up to T_(e2) so asto restrain the engine speed from increasing.

[0112] As understood from the above-mentioned explanation, according tothe present invention, not only the fuel economy but also theacceleration performance can be enhanced. Further, it is possible toprovide a highly flexible control apparatus for a drive system composedof an engine and a transmission.

[0113] Explanation will be made hereinbelow a first embodiment of anengine which is used in the above-mentioned drive system. Referring toFIG. 17, the engine 201 incorporates a piston 202 having a concavecombustion chamber, an intake valve 203, an exhaust valve 204, a fuelinjection valve 205, a spark plug 206, an intake pipe 207 in which anair cleaner 208 is located, and an exhaust pipe 209 in which a catalyticconverter 210 for purifying nitrogen oxide is incorporated. The intakevalve 203 is driven by a low load cam 211 and a high load cam 212. Theexhaust valve 204 is driven by a cam 213. The cam 211 presses a rockerarm 214 while the cam 212 presses a rocker arm 215. In this arrangement,under a low load, a solenoid 216 is energized so as to connect therocker arm 215 with the intake valve 203. The spark plug 206, the fuelinjection valve 205 and solenoids 216, 217 are operated under thecontrol of a control apparatus 218, a position (degree of depression) ofan accelerator pedal 219 detected by a potentiometer 220, a speed of theengine detected by a speed sensor 221 and an air-fuel ratio of exhaustgas detected by an air-fuel ratio sensor 222 are delivered to thecontrol apparatus 218.

[0114] The fuel injection volume injected from the fuel injection valve205 is controlled in accordance with a position of the accelerator pedal219 as shown in FIG. 18. The reason why the fuel injection volume isdecreased when the engine speed is high, is such as to prevent theengine from overrunning as is well-known. Since the ratio between airvolume and fuel volume becomes large so that the mixture is lean whenthe fuel volume is small, the injection timing is retarded up to aposition near to the compression dead-center, as shown in FIG. 19, andaccordingly, fuel is concentrated around the spark plug 206 in order tostabilize the ignition. It is well-known that the injection timing isset to be earlier than the end of an intake stroke so as to promote themixing of fuel and air when the fuel volume is large. It is alsowell-known that the ignition timing is set to be later than theinjection timing, as shown in FIG. 19, and when the fuel volume islarger, it is retarded. The fuel volume, the injection timing and theignition timing shown in FIGS. 18, 19 are controlled by the controlapparatus 218 as mentioned before.

[0115] The shapes of the high load cam 212 and the low load cam 211 areshown in FIG. 20. The low load cam 211 opens the intake valve 203 untila middle point of a compression stroke. On the contrary, the high loadcam 212 which has the same shape as that of a conventional one, closesthe intake valve 203 at the initiation of a compression stroke.Accordingly, the solenoids are switched so that the high load cam 212 iscoupled to the intake valve 203 when the fuel volume is large but thelow load cam 211 is coupled to the intake valve 203 when the fuel volumeis less. Accordingly, the characteristic of air volume can be obtainedas shown in FIG. 21. The exhaust valve 213 is closed at the end of anexhaust stroke, similar to a conventional one. Thus, the air volume issmall as the fuel volume is small, and accordingly, it is possible toprevent the air-fuel ratio from increasing in a range where the fuelvolume is small, as shown in FIG. 22 so as to stabilize the combustioneven though the fuel volume is small. At this time, if the fuel volumeis set to a value (a) as shown in FIG. 22, the air volume is set so asto prevent the air-fuel ratio from being lower than 16 since thenitrogen oxide emission becomes locally maximum around an air-fuel ratioof 16. Accordingly, the closing timing of the intake valve shown in FIG.21 is set so as to allow the air-fuel ratio to satisfy theabove-mentioned condition. Accordingly, as shown in FIG. 22, theincrement of hydrocarbon emission in a range where the fuel volume issmall, can be restrained, and as well, the increment of nitrogen oxideemission can be restrained, as shown in FIG. 22. The air-fuel ratioshown in FIG. 22 is detected by the air-fuel ratio sensor 222, andaccordingly, if air-fuel ratio approaches to a value 16 at the point a,the fuel volume is decreased or the closing timing of the intake valveis advanced so as to correct to the air-fuel ratio. Accordingly, thenitrogen oxide emission is prevented from increasing.

[0116] As shown in FIG. 24, since the closing timing of the intake valveis retarded, the compression stroke is decreased, but the working strokeis not changed. Accordingly, the ratio between expansion and compressionbecomes 2 so that the expansion work is effectively transmitted to thepiston, and as a result, the fuel consumption rate is reduced by 10%.The retardation of the closing timing of the intake-air valve iswell-known in view of a Miller cycle engine. However, it is novel in thecombination of the in-cylinder fuel injection. That is, multipliertechnical effects and advantage can be obtained in the combinationbetween the effect given by a Miller cycle engine which effectively usesan expansion work and the effect of stabilization of fuel due to adecrease in air volume in the engine cylinder.

[0117] In the first embodiment shown in FIG. 17, the closing timing ofthe intake valve 203 is controlled so as to increase the air volume in arange where the fuel volume is large. Alternatively, the superchargepressure may be increased while the closing timing is unchanged in orderto increase the air-volume. In this case, the ratio between expansionand compression becomes 2 even in a range where the fuel volume islarge, and accordingly, the specific fuel consumption can be decreasedas a whole.

[0118] As mentioned above, stable operation can be made at a high excessair factor (high air-fuel ratio) by in-cylinder fuel injection, andfurther, the expansion work can be effectively used by retardation ofthe closing timing of the intake valve. As a result, the fuelconsumption can be greatly reduced while the emission of hydrocarbon andnitrogen oxide is reduced.

[0119] An increase in air volume in the case of the retardation of theclosing timing of the intake valve 203 can be carried out by adjustingthe supercharging pressure or by adjusting the closing timing the intakevalve 203. In the this first embodiment as shown in FIG. 17, it has beenexplained that the closing timing is changed stepwise. However, theclosing timing may also be changed continuously in an easy manner evenwith the use of conventional technology.

[0120] In a conventional in-cylinder fuel injection engine, theinjection timing and the ignition timing have been set under such acondition that the air-volume in an engine cylinder is constant.However, as in this embodiment, it is difficult to sufficiently copewith an increase in the air volume in a range where the fuel volume islarge. If the air volume increases, the concentration of fuel around thespark plug 206 becomes less if the injection timing is unchanged, andaccordingly, the combustion is unstable. In order to avoid thisphenomenon, it is required to control the injection timing and theignition timing, depending upon a variation in the air volume.

[0121] That is, the engine in the first embodiment is essential with theprovision of either one of the following matters.

[0122] (1) the air volume is precisely controlled in accordance with avariation in load (for fuel mass), and the injection timing and theignition timing are precisely controlled in accordance with a variationin load, that is, an air-fuel ratio of exhaust gas is detected by theair-fuel ratio sensor 222 so as to know an error in the air-volume undercontrol, and accordingly, operation for correction is made; and

[0123] (2) the injection timing and the ignition timing are preciselycontrolled in accordance with a variation in load, and further, theinjection timing and the ignition timing are controlled insynchronization with an air-volume control signal such as a controlsignal for the solenoid 216 or 217.

[0124] Referring to FIG. 25, if the air volume has a value G₂, theinjection timing becomes more and more negative as the fuel volume isincreased. That is, since the compression dead center is set to zero,the injection timing is advanced up to a crank angle of −180 deg. thatis, up to the initiation of a compression stroke.

[0125] In a Miller cycle engine, the air volume is lowered to, forexample, a value G₁ if the supercharge pressure is decreased. At thistime, should the fuel injection timing be fixed with respect to the fuelvolume, as a conventional one, the induction timing near astoichiometric air-fuel ratio would be −90 deg. so that the mixing ofair and fuel could not be promoted. On the contrary, according to thepresent invention, the injection timing is advanced to −180 deg. so asto promote the mixing of air and fuel, thereby it is possible tostabilize the combustion.

[0126] Referring to FIG. 26 which is a control flow chart for thepresent embodiment, at step 291, the engine speed is detected, and atstep 292, a position of the accelerator pedal is detected. At step 293,a desired fuel volume is computed from both detected values. That is,the desired fuel is read from a table on which data obtained from achart shown in FIG. 18 is mapped. Referring to FIG. 24, when the airvolume is larger than a value F₁, the air volume is set to a value G₂,but when if it is less than the value F₁, the air volume is set to avalue G₁. At step 293′, the air volume may be set continuously withrespect to the fuel volume. At step 294, an air-fuel ratio is detectedby the air-fuel ratio sensor 222, and at step 295, an actual air volumeis estimated from the detected value. As step 296, the air volume iscorrected in accordance with the estimated air volume by adjusting thesuper-charge pressure, by adjusting the intake pipe pressure with theuse of a throttle valve incorporated in the intake pipe 7, or byadjusting the closing timing of the intake valve. At step 297, with theuse of a table as shown in FIG. 25, which gives an injection timing withrespect a fuel volume, an air volume and an engine speed, the fueltiming is determined, and at step 298, an actual fuel injection iscarried out. At step 299, similar to the determination of the injectiontiming at step 298, the ignition timing is determined with the use of atable which gives an ignition timing with respect to a fuel volume, anair volume and an engine speed, and at step 300, an actual ignition iscarried out.

[0127] In the case of a four-stroke engine, the steps in the flow chartshown in FIG. 26 are carried out once in every two revolutions.Meanwhile in the case of a two-cycle engine it is carried out once inevery one revolution.

[0128] Referring to FIG. 27 which shows an air volume control apparatus,a closing timing control device 301 for the intake valve 203 shown inFIG. 17, is controlled by a solenoid actuator 307. If the closing timingis retarded, the air volume is decreased. If the closing timing isadvanced to the end of an intake stroke, the air volume is increased.Second, a throttle valve 302 is incorporated in the intake pipe 207, andis controlled by a motor actuator 303. If the valve 302 is opened, theair volume is increased, but if it is closed, the air volume isdecreased. Third, a compressor 304 is incorporated, and is driven by amotor or the engine 201 so as to increase the pressure of air in orderto increase the air volume. A bypass valve 305 is actuated by a motor306 so as to be opened, the air volume is decreased. Since the airvolume is changed, depending upon a variation in the atmosphericpressure or the air temperature, output signals from a temperaturesensor 308 and a pressure sensor 309 are delivered to the controlapparatus 218, and accordingly, at step 295 shown in FIG. 26, the airvolume is estimated in accordance with the output signals.

[0129] The maximum air volume is restrained by the stroke volume of theengine 201 and the capacity of the compressor. It is required forincreasing the power or torque of the engine 201 to increase the fuelvolume up to a value at which the air-fuel ratio becomes 11, as shown inFIG. 28, similar to a conventional engine, the density of nitrogen oxide(NOx) exhibits a local maximum value in an air-fuel ratio range from 16to 18, and accordingly, the engine 1 is driven, keeping away from thisrange. As shown in FIG. 29, if the fuel volume is less than a value f₁,the air volume is set to be small so as to set the air-fuel ratio to avalue larger than 18. Meanwhile if the fuel volume is larger than avalue f₂, the air volume is set to be large so as to set the air-fuelratio to a value less than 14.7 (stoichiometric air-fuel ratio). If thefuel volume is between values f₁, f₂, the air volume is continuouslychanged in order to control the air-fuel ratio at a stoichiometricair-fuel ratio. In this range, the nitrogen oxide is purified with theuse of three-way catalyst.

[0130] Referring to FIG. 30 which shows the NOx emission with respect tothe air-fuel ratio, the later the injection timing, the larger theair-fuel ratio which exhibit a peak value of NOx. With the air-fuelratio higher than a mark  on the curve, the combustion of the enginebecomes unstable. Accordingly, the engine is operated on the right sideof the mark . However, if the air-fuel ratio becomes small, theemission of NOx increases, and accordingly, the engine is drivenadjacent to the mark . That is, the setting of the air-fuel ratio withrespect to the injection timing, the setting of the injection timingwith respect to the air-fuel ratio, or the setting of air-fuel ratiowith respect to the fuel volume (refer to FIG. 19) is determined inaccordance with empirical data as shown in FIG. 30. If the injectiontiming is delayed while the air-fuel ratio is unchanged, the emission ofNOx increases. When the air-fuel ratio is decreased with respect to thefuel volume, if the advance control of the injection timing is delayed,the emission of NOx is delayed. However, in this embodiment, since thefuel injection valve 5 is electrically controlled, the injection timingis not delayed, and accordingly, it is possible to prevent the emissionof NOx from increasing.

[0131] As mentioned above, although explanation has been made of thefirst embodiment of the in-cylinder fuel injection engine in which themixture is ignited and burnt by the spark plug, it goes without sayingthat present invention can be applied for a self-ignition engine such asa Diesel-engine. Further, in the first embodiment, although the Millercycle engine having a ratio between compression and working strokes ofless than 1, in which the closing timing of the intake valve isretarded, has been explained as an example, the Miller cycle engine canalso be realized by such an arrangement that the closing timing of theintake valve is advanced, that is, the intake valve is closedintermediate of an intake stroke.

[0132] In this embodiment, the fuel injection timing can be controlledin accordance with the air volume in the engine cylinder in thisembodiment, it is possible to prevent generation of soot, occurrence ofunstable combustion and increasing of NOx.

[0133] Further, since the ratio between compression and working strokesis set to be less than 1 and the sable combustion is attained, thecompression work can be reduced, thereby it is possible to enhance thefuel economy.

[0134] Explanation will be hereinbelow made of a second embodiment of anengine used in the drive system according to the present invention.

[0135] Referring to FIG. 31 which shows an arrangement of an engine andcomponents therearond in the second embodiment, the engine in thisembodiment is a gasoline type four cylinder MiIller cycle engine 310having a cylinder head formed therein with an intake port 313 and anexhaust port 314 which are connected respectively to an intake pipe 320and an exhaust pipe 330. Further, a fuel injection valve 380 and a sparkplug 340 are provided in the cylinder head. Further, an intake valve 315is provided in the intake port 313, and an exhaust valve 316 is providedin the exhaust port 316. The intake pipe 310 is provided therein with athrottle valve 321 for adjusting the flow rate of air flowingtherethrough. Meanwhile, the exhaust pipe 330 is provided therein with acatalytic converter 331 for removing detrimental components from exhaustgas flowing therethrough. A water jacket 318 reserving therein coolingwater is provided around the outer periphery of an engine cylinder 317.The water jacket 318 is connected with a radiator (which is not shown)through the intermediary of a pipe so that the cooling water iscirculated between the water jacket and the radiator.

[0136] An intake and exhaust valve drive mechanism 350 is coupled to theintake valve 315 and the exhaust valve 316. Further, the fuel injectionvalve 380 in each engine cylinder 380 is connected with a fueldistributor (injection timing adjusting means) 360. The throttle valve321 is coupled to an accelerator pedal 323 through the intermediary of awire 322 so as to be operated in association with a degree ofmanipulation of the accelerator pedal 323. The spark plug 340 isconnected-to a spark plug drive circuit 341. The intake and exhaustvalve drive mechanism 350, the fuel distributor 360, the fuel injectionvalve 380 and the ignition plug drive circuit 341 are connected to acontrol unit (ECU) 390 which delivers control signals thereto.

[0137] The intake pipe 320 incorporates an air-flowmeter 319 fordetecting a mass flow rate A of air flowing therethrough. Meanwhile, theexhaust pipe 330 incorporates an exhaust gas thermometer 394 fordetecting a temperature Tg of exhaust gas flowing therethrough. Further,the water jacket 318 incorporates a cooling water thermometer 393 fordetecting a temperature Tw of cooling water flowing therethrough. Thethrottle valve 321 is provided thereto with a throttle opening degreemeter 392 for detecting a degree thereof. The engine has a crankshaft(which is not shown) is provided thereto with a engine speed meter 395for detecting a speed of the engine.

[0138] The air flowmeter 391, the throttle opening degree meter 392, thecooling water thermometer 393 and the exhaust gas thermometer 391 areconnected to the control unit 390 which therefore receives detectionsignals from these meters.

[0139] The control unit 390 is the so-called microcomputer, andincorporates an A/D converter (which is not shown) for converting analogsignals from these meters 391, 392. 393, into digital signals, a ROM(which is also not shown) in which several programs and the like arestored, a CPU (which is not shown) for carrying out several kinds ofcomputation in accordance with the programs stored in the ROM, a RAM(which is not shown) in which detection results from the meters andcomputation results from the CPU are temporarily stored, and the like.This control unit 390 serves as a control means for delivering controlsignals to a fuel injection volume computing means for computing a fuelinjection volume, the distributor 360 as a fuel injection timingadjusting means, the fuel injection valve 380 and the like.

[0140] Referring to FIG. 32, the intake pipe 320 has a straight part inthe vicinity of the intake port 313 of the engine 310. In thisarrangement, the relationship between the radius R of curvature of theintake pipe 320 in the vicinity of the intake port 313 and the pressureloss is shown in FIG. 33. That is, if the radius R of curvature of theintake pipe 320 is 10 cm, the pressure loss of the intake pipe 320 is1×10³ Pa which is a substantially minimum value. Accordingly, eventhough the radius R of curvature is further increased, the pressure losscannot be further decreased substantially. The value 1×10³ Pa given bysetting the radius R of curvature to 10 cm, does not affect the outputpower of the engine 310, substantially. Accordingly, in this embodiment,the radius R of curvature of the intake pipe 320 in the vicinity of theintake port 313 is set to be slightly larger than 10 cm so as to preventair stream from breaking away, in order to reduce the pressure loss asfar as possible. As a result, the air volume charged into the engine 310is increased, and accordingly, the output power of the engine 310 isenhanced.

[0141] In this embodiment, referring to FIG. 34, two intake valves 315and two exhaust valves 316 (only one intake valve 315 a and one exhaustvalve 316 are depicted in the figure) are provided for each enginecylinder. The intake and exhaust valve drive mechanism 350 is adapted tooperate these valves 315, 316 with appropriate timing. The intake andexhaust valve drive mechanism 350 has a cam shaft 351 coupled to thecrankshaft (which is not shown) of the engine 310 through theintermediary of a timing chain, a cam 352 adapted to be rotated inassociation with the rotation of the cam shaft 351, rocker arms 353 a,353 b making contact at one end thereof wit the peripheral surface ofthe cam 352 and at the other end thereof with the stem heads of thevalves 315 a, 316 a, and rocker shafts 354 a, 354 b for swingablysupporting the rocker arms 353 a, 353 b. The rocker arms 353 a, 353 bswing at one end thereof along the peripheral surface of the cam, andaccordingly, press, at the other end thereof, the stems of the valves315 a 316 a which are therefore operated. The lifts and the operationtiming of the valves 315 a, 316 a can be adjusted by changing theprofile of the cam 352. The operation timing of the valves will bedescribed hereinbelow. Although a drive mechanism for the exhaust valves316 is not shown in FIGS. 31 and 34, the basic structure thereof issimilar to the drive mechanism for the intake valves 315 which is shownin FIG. 34.

[0142] In this embodiment, as shown in FIG. 34, the fuel injection valve380 is arranged so as to inject fuel direct into the-cylinder chamber312 of the engine 310. In an intake port fuel injection system as isseen in a general gasoline type engine, fuel sticks to the inner surfaceof the intake pipe 320 and the upper surface of the intake valve 315,and as a result, fuel cannot be fed into the cylinder chamber by adesired volume at a desired time, and accordingly, the combustion in thecylinder chamber possibly becomes unstable. In particular, if the liftof the intake valve 315 is small (that is, less than 1.98 mm), fuelstagnating on the upper surface of the intake valve 315 discretelyenters into the cylinder chamber, causing the combustion to be unstable,and accordingly, the tendency of unstable revolution of the engine ishigh. In view of the foregoing, in this embodiment, the fuel is injecteddirect into the engine cylinder so as to prevent the fuel from stickingto the inner surface of the intake pipe 320 and the upper surface of theintake valve 315. Further, in this embodiment, during an intake strokein a low engine speed range, of two intake valves 315 a, 315 b, the one315 b is temporarily stopped, while the other 315 a is opened so as tocreate a swirl flow in the cylinder chamber 312 in order to promote thecombustion. As a result, as shown in FIG. 35, according to thisembodiment, the engine speed during idle operation is remarkably stable.

[0143] The fuel injection valve 380 is composed of, as shown in FIG. 36,a valve element 386, a position adjuster 387, fuel passages 382, 383, avalve displacement space 385, and a valve casing 381 for housing theabove-mentioned components. The fuel passages 382, 383 have one ene partformed therein with a fuel inlet port (which is not shown) and the otherend part formed therein a fuel jet port 384. The valve displacementspace 385 is formed intermediate of the fuel passages 382, 383, and fuelflows into the valve displacement space 385. That is, a part of thevalve displacement space 385 serves a fuel passage. The passage (whichwill be hereinbelow denoted “valve space outlet side passage”) 383between the valve displacement space 385 and the fuel jet port 384 isformed in a cylindrical shape. The passage (which will be hereinbelowdenoted “valve space inlet port side passage”) 382 is bifurcated intotwo passages 382 a 382 b. One of these passages 382 a, 382 b, which willbe hereinbelow denoted “wide angle atomization passage, is extended,perpendicular to the center axis C of the cylindrical outlet port sidepassage 383, and the other one of them, (which will be hereinbelowdenoted “narrow angle atomization passage“) 382 b is extended in adirection having an obtuse angle to the center axis C of the outlet portside passage 383. The valve element 386 is located in the valvedisplacement space 385 so as to be movable among a valve closingposition where the valve element 386 blocks the valve displacement spaceside port of the valve space outlet side passage 383, a wide angleatomization position (as shown in FIG. 36) where it opens the valvedisplacement space side port of the wide angle atomization passage 382a, but it blocks the valve displacement space side port of the narrowangle atomization passage 382 b and a narrow angle atomization position(as shown in FIG. 37) where it opens the valve displacement space sideport of the wide angle atomization passage 382 a and the valvedisplacement space side port of the narrow angle atomization passage 382b. The position adjuster 387 has a small-sized stepping motor 387 awhich receives a control signal from the ECU 390, and a stopper 387 badapted to be driven by the stepping. motor 387 a. The position adjuster387 locates the valve element 386 at a desired position since thestopper 389 b makes contact with the valve element 385. Specifically,the position adjuster 387 locates the valve element 386 at one of theabove-mentioned valve closing position, wide angle atomization positionand narrow angle atomization position in accordance with a signal fromthe ECU 390.

[0144] When the valve element 386 is located at the valve closingposition, fuel cannot flow from the valve displacement space 385 to theoutlet port side passage 383, and accordingly, no fuel is injected fromthe fuel injection valve 380. When the valve element 386 is located atthe wide angle atomization position, only the wide angle atomizationpassage 382 a which is extended in a direction perpendicular to theoutlet port of the passage 383 is opened. Accordingly, when the fuelcomes out from the wide angle atomization passage 382, the fuel isturned into a swirl flow in the valve displacement space 385, and isjetted from the fuel injection port 384 in a conical shape through theoutlet port side passage 383. Further, when the valve element 386 islocated at the narrow angle atomization position, both wide angleatomization passage 382 a and narrow angle atomization passage 382 b areopened. Since the narrow angle atomization passage 382 b is extended ina direction having an obtuse angle to the outlet port side passage 383,the swirling power of the fuel having come out from the wide angleatomization passage 382 a is decreased. Accordingly, the divergentatomization angle of the fuel jetted from the fuel injection port 384 isnarrower when the valve element 386 is located at the narrow angleatomization position, than when it is located at the wide angleatomization position. Specifically, as shown in FIG. 36, when the valveelement 386 is located at the wide angle atomization position, thedivergent atomization angle of the fuel is 120 deg. while when it islocated at the narrow angle atomization position, the divergentatomization angle is 60 deg.

[0145] Since the engine 310 in this embodiment is a four cylinder engineas mentioned above, the fuel injection valve 380 is provided for each ofthe engine cylinders, that is, four fuel injection valves 380 are intotal provided in the engine 310. The distributor 360 for distributingfuel fed from a fuel tank (which is not shown) by a fuel pump (which isalso not shown) into the fuel injection valves 380, is provided upstreamof the latter, as shown in FIG. 31.

[0146] The distributor 360 comprises a distributor casing 361, a plunger366 located in the casing 361 and adapted to be rotated while beingreciprocated therein, a plunger drive mechanism 370 for moving theplunger 366 while rotating the same, a fuel flow rate adjustingmechanism 368 for adjusting the volume of fuel fed into each of the fuelinjection valves 360 and a fuel injection timing adjusting mechanism 376for adjusting the timing of feeding fuel into each of the fuel injectionvalves 360.

[0147] The distributor casing 361 is formed therein with a plungermoving space 365, a fuel inlet port 362 communicated with the plungermoving space 365, and four fuel outlet ports 363 a, . . . 363 d (referto FIG. 43) respectively communicated with the fuel injection valves 380a, . . . 380 d. The casing fuel inlet port 262 is connected with a fuelpump which is not shown. The plunger 366 is cylindrical, and a main fuelpassage 362 is formed at a position corresponding to the center axis ofthe plunger 366. One end part of the main fuel passage 367 is formedtherein with a fuel inlet port 367 for leading fuel having flown intothe plunger moving space 365 from the fuel inlet port 362 of thedistributor casing 361, into the main fuel passage 367 of the plunger366, and the other end part of the main fuel passage 367 is formedtherein with a fuel discharge port 367 b for returning fuel having flowninto the main fuel passage 367 into a fuel tank (which is not shown). Inan intermediate part of the main fuel passage 37, plunger fuel outletports 367 c, 367 d communicated with the casing fuel outlet ports 363 a,. . . 363 d are formed. As these fuel outlet ports 367 c, 367 d, a firstfuel outlet ports 367 c and a second fuel outlet ports 367 d arepresent, and both ports 367 c, 367 d are symmetric with each other aboutthe center axis of the plunger 366, and are slightly shifted from eachother in the direction in which the center axis of the plunger 366extends.

[0148] The plunger drive mechanism 370 is composed of a cam disc 371fixed to one end part of the plunger 366, a roller 372 making contactwith the outer surface of the cam disc 371 near to the outer peripherythereof, a roller support plate 373 for rotatably supporting the roller372, a cam shaft 374 coupled to the crankshaft of the engine 319 throughthe intermediary of a timing belt of the like, and a connecting rod 375having one end part which is coupled to cam shaft 374 so as to bemovable in the direction of the center axis of the plunger 366, and theother end part fixed to the cam disc 371. The crankshaft of the engine310 and the cam shaft 374 are connected together so that the cam shaft374 is rotated by one revolution as the crankshaft of the engine 310rotates by one revolution. Thus, when the crankshaft of the engine 310rotates by one revolution, the plunger 366 is rotated by two revolutionsabout the center axis thereof by means of the cam shaft 374, theconnecting rod 375 and the cam disc 371. Four convex parts 371 a, 371 b. . . are formed on the outer surface of the cam disc 371 near the outerperiphery thereof. The roller 372 is arranged to make contact with theseconvex parts 371 a, 371 b . . . Accordingly, when the cam shaft 374 isrotated by one revolution, the cam disc 371 and the plunger 366 fixed tothereto are rotated by one revolution which reciprocates them by fourtimes.

[0149] The fuel flow rate adjusting mechanism 368 is composed of a flowrate adjusting ring 368 a which is annular so as to make contact withthe outer periphery of the cylindrical plunger 366 and which isreciprocatable between a position where it blocks the plunger fueldischarge port 367 b and a position where it opens the fuel dischargeport 367 b, a solenoid 368 b for reciprocating the ring 368 a and aconnecting rod 368 c connecting between the solenoid 368 b and the flowrate adjusting ring 368 a.

[0150] The fuel injection timing adjusting mechanism 376 is composed ofan injection timing adjusting ring 377 which is annular so as to makecontact with the outer periphery of the cylindrical plunger 366, andwhich is reciprocatable between a plunger first injection position whereit opens the plunger first fuel outlet port 367 c while it blocks theplunger second fuel outlet port 367 d, and a plunger second injectionposition where it blocks the plunger first fuel outlet port 367 c whileit opens the plunger second fuel outlet port 367 d, and a solenoid 378for reciprocating the ring 377, and a connecting rod 379 connectingbetween the solenoid 378 and the injection timing adjusting ring 377.The injection timing adjusting ring 377 is formed therein withcommunication holes 377 a . . . 377 d which are communicated with thefuel outlet ports 363 a . . . 363 d at the first injection position, asshown in FIG. 43.

[0151] The fuel distributor 360 leads fuel from the casing fuel inletport 362 into the plunger moving space 362 due to the reciprocation ofthe plunger 366 caused by the rotation of the cam shaft 374 whiledischarges fuel having flown into the plunger moving space 365, from theplurality of casing fuel outlet ports 363 a . . . 363 d by way of themain fuel passage 367 of the plunger 366, and the communication holes377 a . . . 377 d in the fuel injection timing ring 377. Which one ofthese casing fuel outlet ports 363 a . . . 363 d discharges fuel isdetermined by a rotating angle of the plunger 366, relative to thecasing 361. The fuel distributor 360 distributes fuel from the fuel tankinto the #1 cylinder fuel injection valve 380 a, the #3 cylinder fuelinjection valve 380 c, the #4 cylinder fuel injection valve 380 d andthe #2 cylinder fuel injection valve 380 b, successively in thementioned order.

[0152] The volumes of fuel discharged from the casing fuel outlet ports363 a, . . . 363 d, are adjusted by the fuel flow rate adjustingmechanism 368. Fuel having flown into the plunger main fuel passage 367from the plunger fuel inlet port 367 a can flow out from the plungerfuel discharge port 367 b, in addition to the plunger fuel outlet ports367 c, . . . 367 d. Accordingly, the fuel discharged from the plungerfuel discharge port 367 b is adjusted by suitably moving the flow rateadjusting ring 368 a of the fuel flow rate adjusting mechanism 366, andaccordingly, the flow rate of fuel discharged, outside of the casing361, from the plunger fuel outlet ports 367 c, 367 d through the casingfuel outlet ports 363 a, . . . 363 d is indirectly adjusted. It is notedthat the fuel discharged from the plunger fuel discharge port 376 b, isreturned into the fuel tank.

[0153] The feed timing of fuel into each of the fuel injection valves380 from the fuel distributor 360 is adjusted by the fuel injectiontiming adjusting mechanism 376. For example, as shown in FIGS. 41 to 44,when the plunger first fuel outlet port 367 c is aligned with the #1cylinder casing fuel outlet port 363 a, and when the fuel injectiontiming adjusting ring 377 is aligned with the first injection position,the plunger first injection port 367 c is communicated with the #1cylinder casing fuel outlet port 363 a through the communication hole377 a in the fuel injection timing adjusting ring 377. Accordingly, thefuel in the main fuel passage 367 of the plunger 366 is fed into the #1cylinder fuel injection valve 380 a through the plunger first fueloutlet port 367 c, the communication hole 377 a in the ring 377, and the#1 cylinder casing fuel outlet port 363 a. Further, even though theplunger first fuel outlet port 367 c is aligned with the #1 cylindercasing fuel outlet port 363 a, if the fuel injection timing adjustingring 377 is located at the second injection position, as shown in FIG.34, the plunger first fuel outlet port 367 c is blocked by the injectiontiming adjusting ring 377 while the plunger second fuel outlet port 367d is opened so that the plunger second fuel outlet port 367 d iscommunicated with the #4 casing fuel outlet port 363 d. Accordingly, thefuel in the main fuel passage 367 of the plunger 366 is fed into the #4cylinder fuel injection valve 380 d by way of the plunger second fueloutlet port 367 d and the #4 cylinder casing fuel outlet port 363 d.Thus, the fuel is not fed into the #1 cylinder fuel injection valve 380a but into the #4 cylinder fuel injection valve 380 d at the timing offeeding fuel into the #1 cylinder fuel injection valve 380 a by movingthe injection timing adjusting ring 377. In other words, the phase ofthe fuel injection can be changed by 180 deg., as shown in FIG. 44 byactuating the fuel injection timing adjusting mechanism 376.

[0154] Next, brief explanation will be hereinbelow made of a Millercycle engine used as the engine 310 in this embodiment.

[0155] A usual four cycle engine has equal compression and workingstrokes. However, the Miller cycle engine 310 has its working strokewhich is longer than its compression stroke, that is, it has a ratiobetween working stroke and compression stroke, which is equal to or lessthan 1, that is, working/compression stroke ratio≧1, in order toincrease the effective work of the engine.

[0156] In this embodiment, the working stroke is set to be longer thanthe compression stroke by controlling the opening and closing timing ofthe intake valve 315.

[0157] Specifically, at first the intake valve 315 is opened while thepiston 311 descends so that air is introduced into the cylinder chamber312 (refer to FIG. 47a, and then, the piston 311 comes to the bottomdead center (refer to FIG. 47b). Thereafter, the intake valve 315 isclosed slightly after piston 311 slightly ascends (refer to FIG. 47c).The compression stroke extends during the period from the time when theintake valve 315 is closed, to the time when the piston comes to the topdead center. The ignition of fuel is carried out just before the timewhen the piston comes to the top dead center (refer to FIG. 47d). Whenthe piston 311 comes up to the top dead center, it is depressed byexplosion of the fuel (refer to FIG. 47e). The working stroke extendsduring the period in which the piston 311 moves from the top dead centerto the bottom dead center (refer to 47 f). The exhaust valve 316 isopened just before the time when the piston 311 comes to the bottom deadcenter. The piston 311 again initiates its ascent so that exhaust gas isdischarged into the exhaust pipe 330 from the cylinder chamber 312 (FIG.47g).

[0158] Thus, in this embodiment, the piston 311 initiates its ascentupon transition between the intake stroke and the compression stroke,but the intake valve is 315 is still opened although the volume of thecylinder chamber is started to be decreased, that is, the intake valve315 is closed, later than that in a conventional engine, andaccordingly, the compression stroke can be set to be shorter than theworking stroke. In other words, the working stroke is set to be longerthan the compression stroke. The control of the opening and closingtiming of the intake valve 315, can be made by changing the profile ofthe cam 52 in the valve drive mechanism 350.

[0159] During the Miller cycle, the compression temperature usuallytends to be lowered so that the evaporation rate of the fuel is loweredsince the compression stroke is short so that the compression ratio issmall. Accordingly, it gives such a disadvantage that the fuel cannot beburnt at a desire air-fuel ratio. Meanwhile, since the compressiontemperature is low, it gives such an advantage that knocking can hardlyoccur.

[0160] Accordingly, in this embodiment, in order to solve theabove-mentioned embodiment, the fuel is injected direct into thecylinder chamber 312, that is, the in-cylinder fuel injection is carriedout. Usually, in the case of fuel injection into the intake port,evaporation of fuel in the intake port part is promoted, andspecifically, the temperature of the intake valve 315 or the cylinderhead is increased in order to evaporate fuel sticking to the innerperipheral surface of the intake pipe, the rear surface of the intakevalve 315 (the surface remote from the surface facing the cylinderchamber) and the like. Accordingly, the intake port injection causeslowering of the charged air volume or lowering of the output power dueto increasing of the temperature. On the contrary, according to thepresent invention, since the fuel is injected direct into the cylinderchamber 312, no fuel sticks to the intake port part, and accordingly, itis not necessary to raise the compression temperature in order toincrease the evaporation rate of fuel. Since no increasing of thecompression temperature is required, a disadvantage inherent to theintake port injection, such as lowering of the output power due to thedecreasing of the charged air volume or possible occurrence of knockingcan be eliminated. That is, in this embodiment, it is possible toenhance the anti-knocking function, and as well to enhance the outputpower of the engine 310 since the charged air volume can be increased.

[0161] Further, since the compression temperature decreases if theclosing of the intake valve 315 is retarded by an angle of greater than30 deg. from the bottom dead center during an intake stroke, theevaporation rate of fuel is lowered so that the combustion in the engine310 becomes unstable. On the contrary, in the this embodiment, eventhough the closing of the vale 315 is retarded by an angle of greaterthan 30 deg. from the bottom dead center during an intake stroke, thestable combustion can be maintained due to the fact as mentioned abovesince the in-cylinder fuel injection is carried out. Accordingly, thereduction of the compression work which is the purpose of the Millercycle engine, can be greatly made. It is noted that the closing of theintake valve 315 with a delay of a crank angle of greater than 30 deg.from the bottom dead-center of an intake stroke, is such that the intakevalve 315 is closed at an angle smaller than an crank angle of 250 deg.from the top dead-center in a compression stroke.

[0162]FIG. 39 shows the relationship between the injection timing andthe density of hydrocarbon in exhaust gas for each fuel divergentatomization angle, which is exhibited by the fuel injection valve 380 inthis embodiment, and which has been explained herein-above. In thisfigure, the injection timing at zero deg. on the abscissa corresponds tothe dead center in a compression stroke.

[0163] As shown in this figure, if the fuel divergent atomization angleis set to 120 deg., the density of hydrocarbon in exhaust gas increasesas the injection timing is retarded (that is, the injection timing ischanged in the direction approaching the top dead center (0 deg.)), andaccordingly, the density of hydrocarbon becomes greatest when theinjection timing is set to a crank angle of about −100 deg. from the topdead center (0 deg.). If the injection timing is retarded further, thedensity of hydrocarbon is contrarily decreased. Further, if the fueldivergent atomization angle is set to 60 deg., the density ofhydrocarbon is lower than that obtained by setting the fuel divergentatomization angle to 120 deg., and is not substantially changed eventhough the injection timing is retarded up to a crankangle of about −40deg. from the top dead center (0 deg.). As mentioned above, the reasonwhy the density of hydrocarbon is higher at 120 deg. of fuel divergentatomization angle than at 60 deg. of fuel divergent atomization angle,is such that the volume of fuel sticking to the wall surface of thecylinder 17 is greater at 120 deg. of fuel divergent atomization anglethan at 60 deg. of fuel divergent atomization angle. Further, after the-injection timing is retarded to about −40 deg. from the dead center (0deg.) in compression stroke, the reason why the density of hydrocarbonis higher at 60 deg. of fuel divergent atomization angle than at 120deg. of fuel divergent atomization angle, is such that the piston 311comes near to the fuel injection valve 380 by retarding the injectiontiming near to the top dead center in compression stroke, andaccordingly the quantity of fuel sticking to the top surface of thepiston 311 is greater at 60 deg. of fuel divergent atomization anglethan at 120 deg. of fuel divergent atomization angle.

[0164] Thus, if the injection timing is advanced, and specifically ifthe fuel is injected into the cylinder before a crankangle of about −40deg. from the top dead center (0 deg.) in compression stroke, the fueldivergent atomization angle is set to 60 deg. Meanwhile, if theinjection timing is retarded, and specifically, if the fuel is injectedinto the cylinder after an crankangle of about −40 deg. from thetop-dead center (0 deg.) in compression stroke, the fuel divergentatomization angle is set to 120 deg. so as to decrease the density ofhydrocarbon in exhaust gas as far as possible.

[0165] By the way, during partial load operation where the fuelinjection volume is less, the mixture around the spark plug 340 is lean,and accordingly, the combustion possibly becomes unstable. Accordingly,the fuel divergent atomization angle is set to 120 deg., as sown in FIG.38, so as to inject a larger volume of fuel toward the spark plug 340during partial load operation. Thus, the concentration of fuel aroundthe spark plug 340 can be maintained to be uniform, and accordingly,stable combustion can be ensured.

[0166] In summary, as shown FIG. 44, during partial load operation, theinjection timing is retarded while the atomization angle is set to 120deg.. Meanwhile during high load operation, the injection timing isadvanced while the atomization angle is set to 60 deg.. Accordingly, thedensity of hydrocarbon in exhaust gas can be lowered, and stablecombustion can be ensured. In this embodiment, in order to carry out theabove-mentioned control, the ECU 390 instructs the fuel injection timingadjusting mechanism 376 in the fuel distributor 360 to retard the fuelinjection timing and instructs the valve position adjuster 357 in thefuel injection valve 380 to set the atomization angle to 120 deg. whenthe fuel injection volume which is determined in accordance with an airflow rate detected by the air flowmeter 391 and an opening degree of thethrottle valve 321 detected by the throttle opening degree meter 392 andthe like, is less than a predetermined value. Further, the ECU 390instructs the fuel injection timing adjusting mechanism in the fueldistributor 360 to advance the fuel injection timing by a crankangle ofabout 180 deg. and instructs the valve position adjuster 357 in the fuelinjection valve 380 to set the atomization angle to 60 deg. when thefuel injection volume determined by the CPU 390 itself exceeds apredetermined value.

[0167] It has been found from experiments made by the applicants, thatthe atomization angle is satisfactorily set to a value larger 100 deg.during partial load operation, and to a value smaller than 90 deg.during high load operation. Further, in this embodiment, when the valveelement is located at the narrow angle atomization position, the valvedisplacement space side port of the wide angle atomization passage 382 aand the valve displacement space side port of the narrow angleatomization passage 382 b are opened. However, it is possible to closethe valve displacement side port of the wide angle atomization passage382 a when the valve displacement space side port of the narrow angleatomization passage is opened.

[0168] By the way, the density of hydrocarbon in exhaust gas iscorrelated to the conversion efficiency of the catalytic converter 331for purifying nitrogen oxide, which will be described later, that is,the conversion efficiency of the catalytic converter 331 for convertingnitrogen oxide into nitrogen is increased as the density of nitrogenoxide increases. In general, rules and regulations given to the densityof nitrogen oxide should be satisfied. Thus, if the density ofhydrocarbon is decreased excessively, the conversion efficiency of thecatalytic converter 331 is lowered so that density of the nitrogen oxidedensity becomes higher, and accordingly, the rules and regulationscannot be sometimes satisfied. Accordingly, in view of the conversionefficiency of the catalytic converter 331, it is required to carry outsuch control that the density of hydrocarbon is lowered in a range wherethe density of nitrogen oxide does not exceed a regulated value.

[0169] Metal ion exchange zeolite catalyst 331 a is located on theengine 310 side of the catalytic converter 331 connected to the exhaustpipe 330, and platinum and alumina group catalyst is located on theexhaust port side thereof. The metal ion exchange zeolite catalyst 331 ahas such characteristics that its low temperature activity is high, butits NOx selective reduction activity is low. Further, the platinum andalumina group catalyst has such that the low temperature activity is lowbut the NOx selective reduction activity is high. Accordingly, in anoperating range where the density of hydrocarbon (HC) is high duringhigh engine speed and high load operation so that, as previouslyexplained with reference to FIG. 40, the density of hydrocarbon (HC) islow during high engine speed, and accordingly, the efficiency ofnitrogen oxide-nitrogen conversion of catalyst tends to be decreased,the platinum alumina group catalyst 318 having a high NOx selectivereduction activity mainly becomes effective if the catalytic environmenttemperature is high. On the contrary, in an operating range where thedensity of HC is high during low engine speed and low load operation sothat, as mentioned above, the efficiency of nitrogen oxide-nitrogenconversion of the catalyst is high, the metal ion-exchange zeolite whichis active even though the catalytic environment temperature is low ismainly effective. It is noted that the reason why the density of HC islow during high engine speed and high load, is such that the oxidativereaction is promoted since the temperature of exhaust gas is high duringa process in which HC is exhausted from the cylinder chamber 312 intothe exhaust pipe 330. Further, the reason why the density of HC is highduring low engine speed and low load operation, oxidative reaction isnot promoted since the temperature of exhaust gas is low, andaccordingly, HC is directly exhausted as it is.

[0170] HC exhausted upon a start of the engine 310 is mainly adsorbed bythe metal ion-exchange zeolite catalyst 331 a. When the temperature ofthe catalytic converter 331 is raised by exhaust gas, HC adsorbed to themetal ion-exchange zeolite catalyst 331 a is separated away, and isoxidized by the platinum alumina group catalyst 331 b. Usual platinumalumina group catalyst is likely to produce nitrogen dioxide N₂O when HCis not oxidized at a low temperature. In order to evade this problem,the platinum alumina group catalyst in this embodiment is added thereinwith palladium or the like in order to enhance the catalytic activity ata low temperature. Further, in order to enhance the conversionefficiency of nitrogen dioxide N₂O during a start of the engine, thefuel injection timing is retarded so as to increase the temperature ofexhaust gas. Alternatively, the ratio between working and compressionstrokes may be changed during Miller cycle, so as to raise thetemperature of exhaust gas.

[0171] Referring to FIG. 36, just after a start of the engine, since thetemperature of exhaust gas is low, HC and NOx are adsorbed onto themetal ion-exchange zeolite catalyst 331 a. During this period, the fuelinjection timing is retarded so as to increase the temperature ofexhaust gas in order to restrain generation of NOx as small as possible.When the temperature of exhaust gas is higher and higher, HC and NOxadsorbed to the metal ion-exchange zeolite catalyst 331 a are graduallyconverted into H₂O, CO₂, N₂. At this time, NOx in the exhaust gas by HCwhich is also adsorbed. When HC adsorbed to the metal ion-exchangezeolite catalyst 331 runs out, the injection timing is advanced so as toincrease the density of HC in exhaust gas. Further, if the temperaturebecomes high, the platinum and alumina group catalyst is mainlyeffective.

[0172] In this embodiment, as mentioned above, control is carried out insuch a way that the injection timing is retarded during partial loadoperation, but the injection timing is advanced during high loadoperation. However, during a start of the engine, control is carried outin such a way that whether the starting of the engine is made or not isrecognized in accordance with a temperature detected by the coolingwater thermometer 393, and if it is starting of the engine, the fuelinjection timing is controlled in accordance with a temperature detectedby the exhaust gas thermometer 394 incorporated to the catalyticconverter 331 in order to enhance the function of the catalyticconverter 331. Specifically, until the temperature detected by thecooling water thermometer 393 rises up to a predetermined temperature,the ECU 390 recognizes that it is warm-up of the engine. In this case,the ECU 390 instructs the fuel injection timing adjusting mechanism 376in the fuel distributor 360 to retard the injection timing until thetemperature detected by the exhaust gas temperature 394 rises up to apredetermined temperature. When the temperature detected by the exhausttemperature thermometer 394 exceeds the predetermined value, the ECU 390instructs the same to advance the injection timing. When the temperaturedetected by the cooling water thermometer 393 exceeds the predeterminedtemperature, the ECU 390 recognizes that the warm-up of the engine hasbeen completed, so as to control the injection timing in accordance witha load.

[0173] As stated above, in the present invention, the in-cylinderinjection is carried out so as to prevent the fuel from sticking to theinner surface of the intake pipe 320 or the upper surface of the intakevalve 315, and accordingly, a desired volume of fuel can be fed into thecylinder chamber 312 at a desired time. Further, since the fuel isprevented from sticking to the inner surface of the intake pipe 320 orthe upper surface of the intake valve 315, the necessity of increasingthe compression temperature for increasing the evaporation rate of fuelcan be eliminated. As a result, it is possible to increase the chargedair volume so as to enhance the output power of the engine and theantiknocking function thereof.

[0174] Further, in this embodiment, during partial load operation, thefuel divergent atomization angle (splay angle) is set to a value greaterthan 100 deg. so as to create a satisfactory mixture around the sparkplug 340. Meanwhile, during high load operation, the fuel divergentatomization angle (splay angle) is set to a value smaller than 90 deg.and the fuel injection timing is advanced to promote the mixing of airand fuel. Thereby it is possible to aim at performing stable combustionover a wide engine operating range.

[0175] Further, in this embodiment, the fuel injection timing and thefuel divergent atomization angle are controlled so as to reduce theexhaust emission of hydrocarbon from the engine 310 itself and thecatalytic converter 331 is efficiently operated. Thus, it is possible toenhance the efficiency of removal of detrimental substance from exhaustgas.

[0176] It is noted that explanation has been made of the embodiment inwhich the present invention is applied to the Miller cycle engine.However, the present invention should not be limited to this embodiment.That is, it goes without saying the present invention can be applied toany other usual engine.

[0177] Next, explanation will be made of variant forms of the fuelinjection valve in this embodiment with reference to FIG. 48 to 50.

[0178] A fuel injection valve 400 in a variant form shown FIG. 48,comprises a valve element 406, a position adjuster 407 for adjusting theposition of the valve element 406, and a valve casing 401 formed thereinwith fuel passages 402, 403 and a valve displacement space 405, andincorporating the valve element 406 and the adjuster 407. The fuelpassages 402, 403 are formed at their one end with fuel inlet ports(which are not shown), and at their the other end with fuel jet ports404. The valve displacement space 405 is formed intermediate of thesefuel passages 402, 403, and the fuel also flows into this valvedisplacement space 405. A plurality of passages 403 (which will behereinbelow denoted “space outlet port side passages”) are formedbetween the valve displacement space 405 and the fuel jet ports 404. Onegroup (which will be hereinbelow denoted “narrow angle atomizationpassages”) of these space outlet port side passages) extend in adirection having an angle of 30 deg. to the injection center axis C, andthe group of the remaining passages 403 a (which will be hereinbelowdenoted “wide angle atomization passages”) extend in a direction havingan angle of 60 deg. to the injection center axis C).

[0179] The valve element 406 comprises a valve end part 406 b adapted toblock the valve displacement space side ports of the wide angleatomization passages 403 a and the valve displacement space side portsof the narrow angle atomization passages 403 b, and a body 406 a havinga front end part formed thereto with the valve end part 406 b. The valvedisplacement space 405 is composed of a valve end part displacementspace 405 b into which only the valve end part 406 b of the valveelement 406 enters, and a body displacement space 405 a into which thebody 406 a of the valve element 406 is fitted. The valve casing 401 isformed therein with a valve seat 401 a at the boundary between the valveend displacement space 405 b and the body displacement space 405 a.

[0180] The valve element 406 is located in the valve displacement space405 so as to be movable among three positions, that is, a valve closingposition where fuel does not flow into the valve end displacement space405 b from the body displacement space 405 (as shown in FIG. 48), a wideangle atomization position where the valve displacement space side portsof the wide angle atomization passages 403 a are opened while the valvedisplacement side ports of the narrow angle atomization passages 403 bare closed (as shown in FIG. 49), and a narrow angle atomizationposition where the valve displacement side ports of the wide angleatomization passages 403 a are closed but the valve displacement sideports of the narrow angle atomization passages 403 b are opened (asshown in FIG. 50).

[0181] The position adjuster 409 comprises a small size stepping motor407 a receiving a signal from the ECU 390, and a stopper 407 b operatedunder the drive of the stepping motor 407 a. In the position adjuster407, the stopper 407 b makes contact with valve element 406 so as tolocate the valve element 406 at a desired position. Specifically, theposition adjuster 407 locates the valve element 406 at one of theabove-mentioned valve closing position, and wide and narrow angleatomization positions in accordance with a signal from the ECU 490.

[0182] Referring to FIG. 48, when the valve element 406 is located atthe valve closing position so as to make contact with the valve seat 401a of the valve casing 401, fuel cannot flow into the valve enddisplacement space 405 b from the body displacement space 405 a, andaccordingly fuel cannot be injected from the valve 400. Further, asshown in FIG. 49, when the valve element 406 is slightly lifted up so asto be located at the wide angle atomization position where the valve endpart 406 b of the valve element 406 blocks the valve displacement spaceside ports of the narrow angle atomization passages 403 b, fuel flowsfrom the valve displacement space 405 into the wide angle atomizationpassages 403 a and is then jetted from the fuel outlet ports 404 a atthe ends of the passages 403 a. At this time, the fuel divergentatomization angle at this time is set to 120 deg. Further, as shown inFIG. 50, the valve element 406 is further lifted up and is located atthe narrow angle atomization position where the valve end part 406 b ofthe valve element 406 blocks only the valve displacement space sideports of the wide angle atomization passages 403 a, the fuel flows fromthe valve displacement space 405 and through the narrow angleatomization passages 403 b, and then are jetted from the fuel outletports 404 b at their ends. At this time, the fuel divergent atomizationangle is set to 60 deg.

[0183] Thus, even with this fuel injection valve 400 in this variantform, the fuel divergent atomization angle can be changed, this fuelinjection valve 400 in this variant form can be used, instead of thefuel injection valve 380 in the aforementioned embodiment, withtechnical effects similar to those obtained by the latter fuel injectionvalve 380.

[0184] By the way, the fuel is jetted from the fuel outlet ports 404 ofthe fuel injection valve 400 in this variant form has a rod-likeatomization shape while the fuel jetted from the single fuel outlet port304 of the fuel injection valve 320 in the aforementioned embodiment hasa conical atomization shape. As shown in FIG. 51, the rod-likeatomization shape gives a density of hydrocarbon which is generallyhigher than that given by the conical shape atomization shape eventhough the air-excessive rate is changed variously. That is, therod-like atomization shape is inferior in fuel distribution, and createsa locally rich mixture. Accordingly, it is preferable to use such fuelinjection valve that the fuel is injected from a single fuel outlet portin a conical atomization shape.

[0185] Referring to FIG. 52 which shows the relationship between theinjection timing of the fuel injection valve which injects fuel in arod-like atomization shape, and the density of hydrocarbon. If theinjection timing is advanced, the quantity of fuel which reaches thewall surface of the cylinder and forms a film before the ignition,increases, and accordingly the density of hydrocarbon becomes high. Asmentioned above, since the density of hydrocarbon relates to theconversion efficiency of the catalytic converter, it is preferable tocontrol the density of hydrocarbon in consideration with this conversionefficiency even though the fuel injection valve 400 having a rod-likeatomization shape as is in this variant form is used.

[0186] Next, explanation will be made of a second variant form of thefuel injection valve with reference to FIG. 53.

[0187] In this variant form, the fuel injection valve 410 has aspherical valve element 412 is used as shown in FIG. 53. This sphericalshape valve element 412, is flexibly connected to a piston 417 throughthe intermediary a pin 416 by a flexible thin connecting rod 415.Accordingly, even though an error such as an eccentricity caused bymachining, is present at a valve seat 411 a or the like of a valvecasing 411, this error can be absorbed by this flexible connecting rod415. When a valve element 419 a of a solenoid valve 419 is lifted up,the pressure of the pressure chamber 418 is lifted up, and the pressurein a pressure chamber 418 is lowered. Accordingly, a piston 419 ispushed up by the force of a spring 414, causing the spherical valveelement 412 to ascend, and the space between the valve seat 411 a andthe spherical valve element 412 is obtained so that the fuel isinjected. The spherical valve element 412, a guide 413, the connectingrod 415 and the piston 417 are movable, and are all small-sized,lightweight and high responsive so that two times of injection per onecycle can be made. When the valve element 419 a of the solenoid valve419 is closed, the pressure of the pressure chamber 418 is increased sothat the piston 417 is depressed, and as a result, the spherical body412 is closed.

[0188] In this arrangement, although the solenoid valve 419 is used foradjusting the pressure in the pressure chamber 418, a laminated typepiezoelectric element can be used, instead of the solenoid valve 418. Asshown in FIG. 54, the laminated type piezoelectric element 420 isprovided in a part of the wall surface forming the pressure chamber 418.Since the fuel flows into the pressure chamber 418, the pressure in thischamber 418 is high. Accordingly, the piston 417 is depressed, andaccordingly, the spherical valve element 412 is pressed against the seat411 a of the casing 411 while the pressure element 420 is pressed. Thus,when the piezoelectric element 420 is pressed, a capacitor 421 ischarged. At an end of a compression stroke, when a switch 422 is closed,the charged capacitor 421 is discharged so that the piezoelectricelement 420 contracts, and accordingly, the pressure in the pressurechamber 112 is lowered. As a result, the piston 417 is slightly raisedso as to slightly lift up the spherical valve element 112, andaccordingly, the fuel is injected.

[0189] Next, explanation will be made of a third variant form of thefuel injection valve with reference to FIGS. 55 to 56.

[0190] As shown in FIG. 55, a fuel induction valve 430 in this variantform comprises a valve casing 431 formed therein with a valvedisplacement space 432 and a solenoid storage part 433, a valve element131 adapted to move in the valve displacement space 432, a spring 436urging the valve element 430 in a valve closing direction, an armature435 fixed to the end part of the valve element 434, a solenoid 437 formoving the valve element 434 together with the armature 435, a solenoiddrive circuit 440 for driving the solenoid 437, and a fuel filter 438for removing foreign matter in fuel flowing into the valve casing 431.The solenoid 437 is energized and deenergized so as to open and closethe valve in order to move the valve element 434. Specifically, when thesolenoid 437 is energized, the valve element 434 is opened, overcomingthe force of the spring. Meanwhile when the solenoid 437 is deenergized,the valve 434 is closed by the force of the spring.

[0191] As shown in FIG. 56, the solenoid drive circuit 440 comprises alower voltage power source 441 and a high voltage power source 442, achange-over switch 443 for applying a voltage from either one of thehigh and low voltage power source to the solenoid 437 and a transistor444 for controlling the current value running through the solenoid 437.When the valve element 434 is lifted (valve opening operation), theswitch 443 is operated so as to a voltage from the high voltage powersource 442 is applied to the solenoid 437. After the valve element 434is completely lifted, if this valve lifting condition is maintained, theswitch 443 is operated so as to apply a voltage from the low voltagepower source 441 to the solenoid 437. Thus, since the high voltage powersource 442 is used for operating the valve element 434, theresponsiveness can be enhanced so that two times of fuel injection canbe made during every cycle. Further, if the valve element 434 is held ata specific position, the low voltage power source 441 is used, and thepower consumption can be decreased while overheating of the solenoid 437can be prevented.

[0192] Although either one of the second and third variant forms of thefuel injection valves cannot change the atomization angle, the flexiblerod 415 which is essential in these variant forms may be used for thefuel injection valves 380, 400 in the first and second embodiments whichcan change the atomization angle. Further, the solenoid 437 may be usedfor displacing the valve element 434 so that the high voltage powersource 442 and the low voltage 441 are changed over by energizing thesolenoid 437.

[0193] Referring to FIG. 57 which shows the cylinder head viewed fromthe cylinder chamber, the injection nozzle 510 is formed therein withtwo jet holes 511, 512 which are adjacent to each other. Electrodes 513,514 of the spark plug are located in an area where the jet holes 511,512 interfere with each other. Flame kernels 517, 518 given by electricdischarge produced between the electrodes 515, 516 of the spark plug aremoved in the direction of the arrow by the atomization 513 and theatomization 514. The velocity of the flow in this area is lower thanthose at the centers of the atomization 513, 514, and accordingly, theflame kernels are not substantially cooled by the atomization 513, 514.When a jet nozzle 517 is closed at a certain time, the flame kernels517, 518 are stopped at the points shown in the figure, and then flamepropagation is started. In this case, as shown FIG. 58, during an intakestroke, main fuel is injected into the cylinder 512 prior to theformation of the flame kernels so as to create a uniform mixture, andthen, ignition fuel is injected at the end of a compression stoke so asto ensure the ignition. Accordingly, it is necessary to inject fuel twotimes per cycle.

[0194] In order to inject fuel two times per cycle, it is required tomodify the distributor 360 stated hereinabove, as shown in FIG. 59. Itis noted in the distributor shown in this figure that like referencenumerals are used to denote like parts to those in the distributor 360so as to avoid redundancy in explanation.

[0195] The distributor 550 in this modified form is the one in which theinjection timing adjusting ring 357 in the above-mentioned distributor360 is modified. In this modified form, the injection timing adjustingring 357 is formed therein with communication holes 357 a, 357 b . . .so that when a casing fuel outlet port 363 a which is one of two casingfuel outlet ports 363 a, 363 d symmetrical with each other about thecenter axis of a plunger is communicated with a plunger first fueloutlet port 367 c, the other casing fuel outlet port 363 d iscommunicated with a plunger second fuel outlet port 367 d. Accordingly,when fuel flows from one of two casing fuel outlet ports which arelocated at positions symmetric with each other about the center axis ofthe plunger, fuel flows also from the other thereof. Specifically, ifthe fuel flows out from a #1 engine cylinder casing fuel outlet port 363a, fuel flows also from a #4 engine cylinder casing fuel outlet port 363d.

[0196] By the way, in the case of a four cylinder engine, when a #1engine cylinder is shifted from an exhaust stroke into an intake stroke,a #4 engine cylinder is shifted from a compression stroke into a workingstroke, as shown in FIG. 44. Meanwhile, when the #1 engine cylinder isshifted from a compression stroke into a working stroke, the #4 enginecylinder is shifted into an exhaust stroke into an intake stroke.Accordingly, as shown in FIG. 58, in such a case that the main fuelinjection and the ignition fuel injection are carried out, when the mainfuel injection is carried out in the #1 engine cylinder, the ignitionfuel injection is carried out in the #4 engine cylinder. Meanwhile, whenthe main fuel injection is carried out in the #4 engine cylinder, theignition fuel injection is carried out in #1 engine cylinder.Accordingly, in this modified form, fuel flows from the #1 enginecylinder casing fuel outlet port 363 a while fuel flows also from the #4engine cylinder fuel outlet port 363 d.

[0197] Incidentally, when main fuel (in a large fuel volume) is fed fromthe casing fuel outlet port 436 s as one of both outlet ports, it isrequired to feed ignition fuel (in a small fuel volume) from the othercasing fuel outlet port 363 d. Accordingly, the ring 457 is formedtherein with the communication holes 457 a, 457 d, . . . so that even ina condition in which the one casing fuel outlet port 363 a and theplunger first fuel outlet port 363 a are completely communicated witheach other, the other casing fuel outlet port 363 d is in halfcommunicated with the plunger second fuel outlet port 367 d.

[0198] It is noted that in order to carry out the main fuel injection inthe #4 engine cylinder, and simultaneously to carry out the ignitionfuel injection in the #1 engine cylinder, the fuel injection timingadjusting ring 457 may be slightly moved rightward.

[0199] If the fuel flows from the casing fuel outlet port 363 d which isone of two casing fuel outlet ports 363 a, 363 d arranged symmetric witheach other about the center axis of the plunger, simultaneously withflowing of fuel from the other casing fuel outlet port 363 a, and if thefuel injection valves have different flow characteristics, either themain fuel injection volume or the ignition fuel injection volume cannotbe fed in their respective desired values due to their different flowcharacteristics. That is, for example, if one of the fuel injectionvalve has a large pressure loss while the other one of them has a smallpressure loss, when the fuel is fed, from one and the same fuel supplysource, simultaneously into the fuel injection valves, fuel having avolume less than the desired value flow from one of the fuel injectionvalves meanwhile fuel having a volume larger than the desired valueflows out from the other one of them. Accordingly, a second modifiedform of the distributor which can feed fuel in desired volumes into thefuel injection valves even though these fuel injection valves havedifferent flow characteristics, will be hereinbelow explained withreference to FIG. 60.

[0200] In this distributor 460 shown in FIG. 60, a plunger second fueloutlet port 467 a is extended in a direction having an angle which isnot 180 deg. to the plunger first fuel outlet port 467 c but is slightlysmaller than 180 deg. around the plunger center axis as a center. It isnoted that the distributor in this second modified form is substantiallythe same as the distributor in the first modified form, excepting theabove-mentioned arrangement. With these plunger fuel outlet ports 467 c,467 d, when fuel is fed into the #4 engine cylinder, no fuel is fed intothe #1 engine cylinder. However, slightly later, the fuel is fed intothe #1 engine cylinder. Accordingly, fuel is not fed simultaneously intothe fuel injection valves from the distributor 460, and therefore, fuelin a substantially desired volume, can be fed into each of the fuelinjection valves even having different flow characteristics.

[0201] Next, explanation will be made of one embodiment of the fuel pumpwith reference to FIGS. 61 to 62.

[0202] As mentioned above, it has been explained that fuel is fed intoeach of the fuel injection valves for engine cylinders from a singlefuel pump by means of the fuel distributor. However, in this embodiment,fuel is fed into the fuel injection valves from respective fuel pumps,that is. the so-called in-line pump will be explained.

[0203] The fuel pump 470 in this embodiment has a pump casing 471, apiston 473 adapted to reciprocate in the pump casing 471, and a pistondrive mechanism 473. This piston drive mechanism 473 comprises a camshaft 474 which is coupled to a crankshaft through the intermediary of atiming belt or the like, an advance cam 475 a and a retardation cam 475b which are fixed to the cam shaft 474 so as to be rotated inassociation with the rotation of the cam shaft 474, and an advance camfollower rod 476 a arranged so as to make contact with the outerperipheral surface of the advance cam 457 a, a retardation cam followerrod 476 b arranged so as to make contract with the outer surface of theretardation cam 475 b, a swingable rod 476 c making contact with an endpart of the piston 472, a support pin 478 for swingably supporting theadvance cam follower rod 476 a, the retardation cam follower rod 476 band the swingable rod 476 c, a timing change-over pin 479 for swingingthe swingable rod 476 c in response to one of the two cam follower rods476 a, 476 b, and a solenoid (which is not shown) for moving the timingchange-over pint 479. The advance cam follower rod 467 a, theretardation cam follower rod 476 b and the swingable rod 467 c arearranged in parallel with one another, and are supported at their oneend part by the support pin 478. The other end part of the swingable rod476 c is formed therein with a timing change-over pin through-hole 477 cthrough which the timing change-over pin 479 pierces, the other end partof the advance cam follower rod 476 a and the other end part of theretardation cam follower rod 476 b are formed respectively thereinchange-over pin fitting parts 477 a, 477 b in which opposite end partsof the timing change-over pin 479 are fitted. The timing change-over pin479 is always inserted in the timing change-over pin through-hole of theswingable rod 476 c, but is fitted, at its either one of the end parts,in either one of the fitting part 477 a of the advance cam follower rod467 a and the retardation cam follower rod 467 b in accordance with itsown position.

[0204] Explanation will be hereinbelow made of the operation of the fuelpump 470.

[0205] When the cam shaft 474 is rotated in association with therotation of the crankshaft, the advance cam 475 a and the retardationcam 475 b fixed to the cam shaft 474 are rotated. In association withthe rotation of these cams 475 a, 475 b, the cam follower rods 476 a,476 b making contact with the outer peripheral surfaces of these cams475 a, 475 b are swung around the support pin 478 as a center inaccordance with shapes of the cams with which they make contact. If thetiming change-over pin 479 is fitted in the fitting part 477 a of theadvance cam follower rod 476 a at this time, the swingable rod 467 c isalso swung in association with the swinging of the advance cam followerrod 476 c. Alternatively, if the timing change-over pin 479 is fitted inthe fitting part 477 b of the retardation cam follower rod 476 b at thistime, the swingable rod 467 c is swung in association with the swingingof the retardation cam follower rod 476 b. Accordingly, the piston 472is reciprocated in association with the swinging of the swingable rod476 c.

[0206] As mentioned above, when the solenoid is energized in response tothe ECU 90 so as to move the timing change-over pin 479, the piston canbe actuated, following the actuation of one of the cams 475 a, 475 b,and accordingly, the fuel injection timing can be changed as shown inFIG. 63.

[0207] The fuel injection timing and the fuel divergent atomizationangle are uniformly related to the density of hydrocarbon. Further, theconversion efficiency of the catalyst and the density of hydrocarbon inexhaust gas are also uniformly related to each other. Accordingly, ifthese relationships are previously examined, and if the fuel injectiontiming and the fuel divergent atomization angle are controlled with theuse of the fuel injection valves according to the present invention,detrimental substances in exhaust gas can be efficiently removed.

[0208] Further, during partial load operation, the fuel divergentatomization angle is widened so as to direct fuel into the electrodes ofthe spark plug from the fuel injection valve so as to create asatisfactory mixture around the electrodes of the spark plug. Meanwhile,during high load operation, the fuel divergent atomization angle isnarrowed while the fuel injection timing is advanced so as to promotethe mixing of air and fuel. Thus, it is possible to aim at carrying outstable combustion over a wide range.

What is claimed is:
 1. A control apparatus for a drive system composedof an engine and a transmission, characterized by a computing means forcontrolling the drive system in relation to a ratio between wheel speedand engine speed and an air-fuel ratio in accordance with a desiredwheel torque and a vehicle speed.
 2. A control apparatus for a drivesystem composed of an engine and a transmission, characterized by acomputing means for controlling the drive system in relation to a ratiobetween wheel speed and engine speed and an intake valve closing anglein accordance with a desired wheel torque and a vehicle speed.
 3. Acontrol apparatus for a drive system composed of an engine and atransmission, characterized by a computing means for controlling thedrive system in relation to a ratio between wheel speed and engine speedand a supercharged pressure in accordance with a desired wheel torqueand a vehicle speed.
 4. A control apparatus for a drive system composedof an engine and a transmission, characterized by a computing means forcontrolling the drive system in relation to a ratio between wheel speedand engine speed and a ratio between working and compression strokes inaccordance with a desired wheel torque and a vehicle speed.
 5. A controlapparatus as set forth any one of claims 1 to 4, characterized in thatsaid computing means carries out such as to optimize fuel economy andacceleration performance with the use of an operation chart.
 6. Acontrol apparatus as set forth any one of claims 1 to 4, characterizedin that said computing means carries out computation in accordance witha taste of a driver relating to acceleration or fuel economy, or a driveenvironment surrounding a vehicle.
 7. A method of controlling aninternal combustion engine in which a volume of fuel injected from afuel injection valve having a jet port in a combustion chamber of saidinternal combustion engine, and injection timing are controlled, andopening and closing timing of an intake valve is also controlled,characterized in that said volume of fuel injected from said fuelinjection valve and said injection timing are controlled in accordancewith a variation in a volume of air to be burnt in said combustionchamber.
 8. A method of controlling an internal combustion engine inwhich a volume of fuel injected from a fuel injection valve having a jetport in a combustion chamber of said internal combustion engine, andinjection timing are controlled, and opening and closing timing of anintake valve is also controlled, characterized in that said volume offuel injected from said fuel injection valve and said injection timingare controlled in accordance with a position of an accelerator pedal. 9.A control apparatus for an internal combustion engine comprising: a fuelinjection timing control means for controlling a volume of fuel injectedfrom a fuel injection valve having a jet port in a combustion engine ofsaid internal combustion engine, and fuel injection timing, and anintake valve opening and closing control means for controlling openingand closing timing of an intake valve, characterized in that said fuelinjection timing control means controls said volume of fuel and saidfuel injection timing in accordance with a variation in a volume of airto be burnt in said combustion chamber.
 10. A control apparatus for aninternal combustion engine comprising: a fuel injection timing controlmeans for controlling a volume of fuel injected from a fuel injectionvalve having a jet port in a combustion engine of said internalcombustion engine, and fuel injection timing, and an intake valveopening and closing control means for controlling opening and closingtiming of an intake valve, characterized in that said fuel injectiontiming control means controls said volume of fuel and said fuelinjection timing in accordance with a position of an accelerator pedal.11. A fuel injection valve for injecting fuel into an engine cylinder inan internal combustion engine, comprising: a valve casing definedtherein a fuel passage having a one end part formed therein with a fuelinlet port and the other end part formed therein with a fuel jet port,and a valve displacement space formed intermediate of said fuel passage,a valve element displaceably incorporated in said valve displacementspace in said valve casing, and a valve position adjusting means foradjusting a position of said valve element in said valve displacementspace, characterized in that said fuel passage formed in said casingcomprises a narrow angle atomization passage by which a divergentatomization angle of fuel atomized when flowing therethrough is set to apredetermined specific angle, and a wide angle atomization angle of fuelatomized when flowing therethrough is set to be larger than saidpredetermined specific angle, said valve displacement space is formed insaid valve casing so that said valve element is movable among a narrowangle atomization position where said fuel flows in said narrow angleatomization passage but does not flow in said wide angle atomizationpassage, a wide angle atomization position where said fuel flows throughsaid wide angle atomization passage, and a valve closing position wheresaid fuel does not flow both wide and narrow angle atomization passages,and said valve position adjusting means displaces said valve elementamong said narrow angle atomization position, said wide angleatomization position and said valve closing position.
 12. A fuelinjection valve as set forth in claim 11, wherein said fuel passagecomprises a space inlet port side passage for leading said fuel fromsaid fuel inlet port into said valve displacement space, and a spaceoutlet port side passage for leading said fuel from said valvedisplacement space into said fuel outlet port, said space outlet portside passage is formed in a cylindrical shape around an injection centeraxis extending in a predetermined direction, as a center axis, saidspace inlet port side passage is bifurcated into two branch passagesjust upstream of said valve displacement space, one of said branchpassages being formed such that fuel flowing into said valvedisplacement space therefrom swirls around said injection center axis,and the other one thereof being formed such that fuel flowing into saidvalve displacement space therefrom weakens a swirling force of fuelwhich flows into said valve displacement space from said one branchpassage so as to swirl in the valve displacement space, and said onebranch passage defines said wide angle atomization passage, and said theother branch passage defines said narrow angle atomization passage. 13.A fuel injection valve as set forth in claim 11, wherein said fuelpassage comprises a space inlet port side passage for leading said fuelfrom said fuel inlet port into said valve displacement space, and aplurality of space outlet port side passages for leading said fuel fromsaid valve displacement space into said fuel outlet port, said aplurality of space outlet port side passages have said fuel outletports, respectively, at their ends, A part of said plurality of spaceoutlet port side passages extends in a direction having a predeterminedspecific angle to an injection center axis which extends in apredetermined direction, and the remaining part thereof extends in adirection having an angle larger than said specific angle, to saidinjection center axis, and said part of said plurality of space outletside passages, define said narrow angle atomization angle, and theremaining part thereof define said wide angle atomization passages. 14.An internal combustion engine apparatus comprises: a four-cycle engineincluding a cylinder, a piston reciprocating in said cylinder, and aspark plug for sparking in said cylinder a fuel injection valve as seforth in claim 1, 2 or 3, a fuel injection timing adjusting means foradjusting a timing with which fuel is injected from said-fuel injectionvalve, a fuel injection volume computing means for computing a volume offuel injected from said fuel injection valve, and a control means forinstructing said fuel injection timing adjusting means to changeinjection timing of fuel from said fuel injection valve, and instructingsaid valve position adjusting means of said fuel injection valve tochange the position of said valve element in dependence upon whethersaid fuel injection volume obtained by said fuel injection volumecomputing means becomes a predetermined value or not,; characterized,said fuel injection valve feeds fuel direct into said cylinder, and isprovided in said cylinder so that fuel which is injected when said valveelement is located at said wide angle atomization position is directedto an electrode of said spark plug.
 15. An internal combustion engineapparatus as set forth in claim 14, wherein said four-cycle engine inwhich said piston reciprocates so as to repeat an intake stroke, acompression stroke, a working stroke and an exhaust stroke, has a Millercycle performing means for making said working stroke larger than saidcompression stroke.
 16. An internal combustion engine apparatus as setforth in claim 14 or 15, further comprising catalyst for removingdetrimental substance from exhaust gas exhausted from said four-cycleengine, wherein said catalyst is composed of metal ion-exchange zeolitecatalyst and platinum alumina group catalyst.